Pivoting engine mount for a model vehicle

ABSTRACT

A mount for a model vehicle engine is provided, having a hinge configured to be secured to a vehicle chassis, a vehicle engine having a crank shaft for driving a vehicle, the engine being secured to the hinge for pivotal movement toward and away from a shaft coupled directly or indirectly to a wheel for driving a vehicle.

RELATED APPLICATIONS

This application claims the benefit of priority under 35 U.S.C. 120 ofprovisional patent application Serial No. 60/669,664 entitled “MOTOROPERATED VEHICLE,” filed on Apr. 7, 2005, and of the patent applicationof Brent W. Byers, Ser. No. 11/102,008 entitled “A MODEL VEHICLESUSPENSION CONTROL LINK,” filed on Apr. 7, 2005 and previouslyincorporated as an Appendix of the aforementioned provisional patentapplication, the contents of which are hereby incorporated by referencein full as if fully set forth herein. This application is also acontinuation-in-part of U.S. design patent application Ser. No.29/227,305 entitled “VEHICLE MOUNTED COIL SPRING AND SHOCK ASSEMBLY”filed on Apr. 7, 2005, the contents of which are hereby incorporated byreference in full as if fully set forth herein.

FIELD OF THE INVENTION

The present invention relates to vehicle design and has particularapplication is the design of remote control and model vehicles.

APPENDICES

Also attached and made a part of this application are Appendices A-C.Appendix A is a document entitled “Model 5310 Revo Owner's Manual” anddescribes in further detail the construction and operation of anembodiment of the invention. Appendix B are documents entitled “TraxxasService and Support Guide” and “Revo Part List,” which describe infurther detail the construction and assembly of components employed inan embodiment of the invention. Appendix C is a document entitled “RevoSuspension Claims,” which describes “progressiveness” in further detailas related to motion ratios and the change in motion ratio.

These Appendices are incorporated by reference in this application intheir entireties to the same extent as if fully set forth herein.

BACKGROUND OF THE INVENTION

Vehicles in a variety of styles and sizes have been made for many years.However, despite improvements in design of vehicles over the years,vehicles remain unduly expensive to construct, expensive to maintain.Furthermore, vehicles, in particular, remotely controlled vehicles suchas models and other reduced-size vehicles, do not have optimum handlingcharacteristics and are unduly difficult to adjust to obtain optimumhandling characteristics under different driving conditions.

Accordingly, it is an object of the present invention to overcome theforegoing limitations of the prior art.

SUMMARY OF THE INVENTION

These and other objects and advantages are achieved in accordance withan embodiment of the present invention, wherein a mount for a modelvehicle engine is provided, comprising a hinge configured to be securedto a vehicle chassis, a vehicle engine having a crank shaft for drivinga vehicle, the engine being secured to the hinge for pivotal movementtoward and away from a shaft coupled directly or indirectly to a wheelfor driving a vehicle.

BRIEF DESCRIPTION OF THE DRAWINGS

For a more complete understanding of the present invention, and theadvantages thereof, reference is now made to the following descriptionstaken in conjunction with the accompanying drawings, in which:

FIG. 1 is in isometric view of a portion of the vehicle showing anengine mount supporting an engine on a chassis, wherein the engine iscoupled to a transmission assembly;

FIGS. 2A through E illustrate an engine mount allowing adjustment of thecenter distance between the engine crankshaft and the transmission inputshaft or engagement and disengagement of a vehicle engine with atransmission;

FIGS. 3A and B are respectively a partial section view, taken along thesection lines of FIG. 2B, and in isometric view of a partial sectionview;

FIGS. 4A through C are top, front elevation and side views of thatportion of the vehicle chassis on which the engine and transmission aremounted;

FIG. 5 is a partial section view of the engine and any amount, takenalong the section lines of FIG. 4B;

FIGS. 6A through D are isometric, front elevation, side, and top viewsof an engine and throttle link assembly of a vehicle;

FIG. 7 is a detail perspective view of a portion of the throttle linkassembly illustrated in FIG. 6A;

FIG. 8 is a partial section view of the throttle link assembly, takenalong the section lines of FIG. 6C;

FIGS. 9A through D are perspective, front elevation, side and top viewsof a front portion of the vehicle, on which is mounted a bumperassembly;

FIG. 9E is a section view, taken along the section line of FIG. 9C;

FIG. 10 is a perspective view of a vehicle chassis with the body shellremoved;

FIG. 11 is a sectional view of the vehicle chassis of FIG. 10, takenthrough the portion of the vehicle chassis including the fuel tank,filler cap and finger pull tab, with the cap open, along the line 10-10;

FIG. 12 is a perspective sectional view of a vehicle chassis, with thebody shell installed, taken through the portion of the vehicle chassisincluding the fuel tank, filler cap and finger pull tab, with the capopen, and showing one half of the opening through with the finger pulltab can pass when the body shell is installed or removed;

FIG. 13A is a plan view of the fuel tank, filler cap and finger pulltab, with the cap open;

FIG. 13B is a side view of the fuel tank, filler cap and finger pulltab, as viewed from the rear of the vehicle, with the cap open;

FIG. 13C is a perspective view of the fuel tank, filler cap and fingerpull tab, with the cap open;

FIG. 13D is a side plan view of the fuel tank, filler cap and fingerpull tab, as viewed from the right side of the vehicle, with the capopen;

FIG. 14 is a partially sectional view of the fuel tank, filler cap andfinger pull tab, taken along the line 14-14, with the cap open;

FIG. 15 is a perspective sectional view of a vehicle chassis, with thebody shell installed, showing the cap opened;

FIG. 16 is a plan view of a vehicle chassis with the body shell andsuspension components removed;

FIG. 17 is a sectional view of the vehicle chassis of FIG. 16, takenalong the line 16-16, with a detail circle K around the secured doublelooped fuel line in accordance with an embodiment of the presentinvention;

FIG. 18 is a perspective view of the vehicle chassis of FIGS. 16 and 17,showing the secured double looped fuel line;

FIG. 19A is a detailed perspective view showing the secured doublelooped fuel line;

FIG. 19B is a detailed cross-sectional view taken within the detailcircle of FIG. 17, showing a cross-section of the secured double loopedfuel line as secured in its chassis mount;

FIGS. 20A through C are front, side in perspective views of a slipperclutch assembly for use in a vehicle;

FIGS. 21A and B are exploded in perspective views of the slipper clutchassembly;

FIG. 22 is a section view, taken along the section lines of FIG. 20A;

FIG. 23 is an enlarged detail illustration of a portion of FIG. 22;

FIG. 24 is a partial section view of the slipper clutch assembly;

FIG. 25A is an axial view, looking along the axis of the brake disk fromthe outboard side, of a brake pad support assembly in accordance withone embodiment of the present invention;

FIG. 25B is a side view of the brake pad support assembly depicted inFIG. 25A;

FIG. 25C is a plan view of the brake pad support assembly depicted inFIG. 25A;

FIG. 25D is a perspective view of the brake pad support assemblydepicted in FIG. 25A, as viewed from the outboard side;

FIG. 26A is a sectional view of the brake pad support assembly depictedin FIG. 25A, taken along the line 25A-25A of FIG. 25A;

FIG. 26B is a sectional perspective view of the brake pad supportassembly depicted in FIG. 25D, taken along the line 25D-25D of FIG. 25D;

FIG. 27 is an exploded perspective view of an embodiment of the brakepad support assembly and base, as viewed from the outboard side;

FIG. 28 is an exploded perspective view of an embodiment of the brakepad support assembly and base, as viewed from the inboard side;

FIGS. 29A through D are rear elevation, side, top and perspective viewsof a front bulkhead assembly and suspension arm assembly of the vehicle;

FIGS. 30A through D are front elevation, side, top and perspective viewsof a telescoping drive shaft of the vehicle;

FIGS. 31A and B are section and perspective section views, taken alongthe section lines 31-31 of FIG. 30A, of the telescoping drive shaft;

FIGS. 32A and B are section and perspective section views, taken alongthe section lines 32-32 of FIG. 30A, of the telescoping drive shaft;

FIGS. 33A through D are rear elevation, side, top and perspective viewsillustrating coupling of the drive shaft to an axle assembly supportinga wheel of the vehicle;

FIG. 34 is a section view, taken along the section lines 34-34 of FIG.33C, illustrating coupling of the drive shaft to an axle assemblysupporting a wheel of the vehicle;

FIG. 35 is a perspective section view, taken along the section lines35-35 of FIG. 33C, illustrating coupling of the drive shaft to an axleassembly supporting a wheel of the vehicle;

FIG. 36 is a section view substantially bisecting the ball joint andaxle carrier assemblies of the vehicle;

FIG. 37 is a side view of the axle carrier shown in FIG. 36;

FIG. 38 is a perspective exploded view of the axle carrier showing asealing boot secured to the carrier;

FIGS. 39A through C are front elevation, side and top views of the axlecarrier shown in FIG. 38;

FIG. 40A is view of the front portion of the vehicle, with the chassisremoved for clarity, showing the dual servos and center dual armsteering arm, viewed from underneath;

FIG. 40B is view of the front portion of the vehicle, with the chassisremoved for clarity, showing the dual servos and center dual armsteering arm, viewed from the front end of the vehicle;

FIG. 40C is view of the front portion of the vehicle, with the chassisremoved for clarity, showing the left side front wheel and left sideservo and the center dual arm steering arm, viewed from the left side ofthe vehicle;

FIG. 40D is a perspective view of the front portion of the vehicle, withthe chassis removed for clarity, showing the dual servos and center dualarm steering arm, viewed from underneath the left side of the vehicle;

FIG. 41A is an exploded perspective view of the components of the dualservos and center dual arm steering arm assembly, as viewed from abovethe vehicle;

FIG. 41B is an exploded perspective view of the components of the dualservos and center dual arm steering arm assembly, as viewed from belowthe vehicle;

FIG. 42 is a perspective view of the dual servos and center dual armsteering arm assembly, with the other components of the front end of thevehicle removed for clarity, viewed from the rear left side of thevehicle;

FIG. 43A is a plan view of a steering servo mounted on the right side ofthe chassis;

FIG. 43B is a side view of a steering servo mounted on the right side ofthe chassis;

FIG. 43C is a perspective view of a steering servo mounted on the rightside of the chassis;

FIG. 43D is an end view of a steering servo mounted on the right side ofthe chassis, viewed from the front of the vehicle;

FIG. 44 is a sectional view of the mounted steering servo of FIG. 42A,taken along the line 41A-41A;

FIG. 45 is a perspective view of a steering servo mounted on the rightside of the chassis, and shows a front one of the mounting brackets;

FIG. 46 is an exploded perspective view of a steering servo, front andrear mounting brackets, and the portion of the chassis to which thesteering servo is mounted;

FIGS. 47A and B are side and top plan views showing the layout ofvarious components supported by the vehicle chassis;

FIG. 48 is a perspective view of a vehicle chassis alone;

FIGS. 49A through D are side, front, top and perspective views of thevehicle chassis supporting certain components of a vehicle;

FIGS. 50A and B are section and perspective section views, taken alongsection lines of FIG. 49C, illustrating the shape of the chassis andrelative location of certain components supported by the chassis;

FIGS. 51A and B are section and perspective section views, taken alongsection lines of FIG. 49C, illustrating the shape of the chassis andrelative location of certain components supported by the chassis;

FIG. 52 he is a section view, taken along section lines of FIG. 49C,illustrating the shape of the chassis and relative location of certaincomponents supported by the chassis;

FIG. 53, depicts a perspective view of the front suspension assembly forthe left front wheel;

FIGS. 54A-E show detailed views of the axle carrier, pin and pivot linkwith various predetermined combinations of ring-shaped spacers; and

FIG. 55 is a table depicting an example of five different positioningsof the pivot link for different combinations of caster angle and rollcenter settings, employing a thick spacer and a thin spacer in differentconfiguration, as well as a standard configuration employing a tallcenter hollow ball type pivot link.

FIG. 56 is an exploded perspective view of the front left suspensionassembly of the vehicle;

FIGS. 57A through D are front elevation, side, top and perspective viewsof the front left suspension assembly of the vehicle in a full bumpposition;

FIGS. 58A through D are front elevation, side, top and perspective viewsof the front left suspension assembly of the vehicle in a full droopposition;

FIG. 59 is a dimensioned front elevation of the front left suspensionassembly of the vehicle, shown at ride height;

FIG. 60 is a dimensioned rear elevation of the rear left suspensionassembly of the vehicle, shown at ride height;

FIG. 61 is a dimensioned top view of the chassis of the vehicle showingthe front and rear left suspension assemblies of the vehicle;

FIGS. 62A and B are top and side views of a rocker arm employed in arear suspension assembly of the vehicle;

FIGS. 63A and B are top and side views of a rocker arm employed in thefront suspension assembly of the vehicle; and

FIG. 64 is top view of a portion of the front left suspension assemblyof the vehicle showing the damper and rocker arm employed therein.

DETAILED DESCRIPTION

FIG. 1 illustrates a vehicle engine 500 supported by an engine mount 510(partially shown) on the vehicle chassis 300. The engine 500 drive shaft512 rotates a clutch bell 514 and drive gear 516 assembly that iscoupled via a spur gear 518 to a transmission assembly 520. The enginemount 510 is configured to allow generally vertical movement, shown bythe arrows 522, to accommodate drive and spur gears 516, 518 ofdifferent sizes or to allow engagement and disengagement of a vehicleengine with a transmission. Such gear mesh adjustment, in a generallyvertical direction, reduces horizontal space needed on the chassis 300and accommodates the multi-level design of the chassis 300.

Referring now to FIGS. 1, 2A through E, 3A and B and 4A through C, theadjustable engine mount 510 is shown in more detail. The engine mount510 comprises a front support 524, a middle support 526 and a rearsupport 528. The supports 524, 526 and 528 are preferably manufacturedfrom cast aluminum; however, other suitable materials having therequired strength and temperature resistance would also be suitable. Thefront and rear supports 524, 528 are generally rib-shaped and aresecured on the chassis 300 by outboard flanges 530 and inboard flanges532. Bolts 534 are inserted into threaded apertures 535 formed in theflanges 530, 532 from and through the bottom of the chassis 300. Themiddle support 526 is pivotally mounted to the front and rear supports524, 528 by a pivot bolt 536 extending through a hinge aperture 538 of amiddle support 526 and aligned apertures 540, 542 through the front andrear supports 524, 528 respectively. The pivot bolt 536 comprises athreaded end 554, but preferably has a smooth surface that extendsthrough the hinge aperture 538. The threaded end 554 secures the pivotbolt 536 to a threaded shank 546 extending laterally from and inalignment with the aperture 540 of the front support 524. The smoothsurface of the pivot bolt 536 reduces friction, thereby facilitatingpivoting of the middle support 526 between the front and rear supports524, 528.

The middle support 526 includes a pivot arm 547 extending generallydownwardly and inboard from the remainder of the support 526. The pivotarm 547 positions the hinge aperture 538 so as to impart a horizontalcomponent to the pivotal movement of the engine 500 when the middlesupport 526 is pivoted from the lowest to the uppermost position. Therotational axis of the drive gear 516 is offset in the outboarddirection from the rotational axis of the spur gear 518. Thus, thehorizontal component of movement of the engine 500 as the middle support526 pivots upwardly, moves the drive gear 516 axis more directly towardthe spur gear 518 axis than would otherwise be the case, facilitatingmeshing of the gears with reduced interference. The pivot arm 547 alsopositions the hinge aperture 538 inboard, to impart greater movement ofthe engine 500 as the middle support 526 is pivoted. The pivot arm 547is formed from a plurality of structural ribs 549, to reduce the weightof the middle support 526.

Setting of the position of the engine mount 510 is accomplished by anadjustment bolt 546, which extends through an aperture 548, anadjustment slot 550 and an aperture 552, through the respective rearsupport 528, middle support 526 and front support 524. The adjustmentslot 550 is located near the outboard end of the middle support 526, forease of access and clearance from the engine 500. A lock washer (notshown) is positioned over the adjustment bolt 546, between the surfacesof the rear and middle supports 528, 526 and between the services of themiddle and front supports 526, 524, to secure the surfaces againstrelative movement when the adjustment bolt 546 is tightened. Theadjustment bolt 546 comprises a threaded end 554, but preferably has asmooth surface that extends through the adjustment slot 550. Thethreaded end 554 secures the adjustment bolt 546 to a threaded shank 556extending laterally from and in alignment with the aperture 552 of thefront support 524. The smooth surface of the adjustment bolt 546 reducesfriction, thereby facilitating pivoting of the middle support 526between the front and rear supports 524, 528.

The engine 500 is supported by inboard and outboard engine supportsurfaces 558, 560 formed on the engine mount 510 middle support 526.Threaded engine fastening bores 562 are formed through the supportsurfaces 558, 560, to receive threaded engine fastening bolts 564. Thefastening bolts 564 are tightened into the engine fastening bores 562and through outboard and inboard flanges 566 extending laterally fromthe engine 500, to secure the engine 500 to the pivotable middle support526 of the engine mount 510. The engine mount 510 is generally U-shapedbetween the engine support surfaces 558, 560, to receive the lower endof the engine 500.

In use, the engine mount 510 may be employed to position the engine 500drive gear 516 toward and away from the spur gear 518. The adjustmentbolt 546 is loosened, allowing the outboard end of the middle support526 of the engine mount 510 to be pivoted to a desired position, aboutthe pivot bolt 536, parting the drive gear 516 and the spur gear 518.The middle support 526 acts as a hinge relative to the chassis 300 andthe transmission assembly 520, which is fixed to the chassis 300. Therange of pivotal movement of the middle support 526 is determined by thelength of the adjustment slot 550. The length of the adjustment slot 550is determined, primarily based on the variety of teeth or sizes of thedrive gear 516 and spur gear 518. The centerline of the adjustment slot550 substantially tracks a constant radius from the pivot boltcenterline 536, to allow pivotal movement of the middle support 526without substantial interference between the surfaces of the adjustmentbolt 546 and the adjustment slot 550. Once substitution of a differentsized drive gear 516 or spur gear 518 is made, or other modifications ormaintenance is completed, the engine 500 is pivoted upwardly to mesh thedrive gear 516 and spur gear 518, connecting the engine 500 to thetransmission assembly 520. The adjustment bolt 546 is then tightened,securing the middle support 526 in the desired position for operation ofthe vehicle engine 500 and transmission assembly 520.

Referring now to FIGS. 6A through D, 7 and 8 a throttle link assembly600 is shown that accommodates vertical movement of the engine 500 bythe engine mount 510 without being uncoupled from the engine 500. Thethrottle link assembly 600 is mounted to the middle support 526 of theengine mount 510, for movement with the engine 500 and the throttle arm602 extending downwardly from the engine throttle 604. The middlesupport 526 includes a throttle link support surface 606 (shown in FIGS.1 through 3B) extending towards the front of the vehicle. The throttlelink support surface 606 includes a threaded aperture into which isthreaded a throttle link bolt 608, securing the throttle link assembly600 for pivotal movement about an axis generally perpendicular to thethrottle link support surface 606.

The throttle link assembly 600 includes a bell crank 610 secured forpivotal movement about the bolt 608, to actuate the throttle arm 602 inresponse to actuation of a servo-link 612. The bell crank 610 includes acentral cylindrical shaft 614, through which the bolt 608 extends. Thebell crank 610 pivots about bolt 608. A servo-link arm 616 and athrottle actuation arm 618 extend in substantially perpendiculardirections from bell crank 610. The servo-link 612 and the throttle arm602 are both pivotally connected to the servo-link arm 616 and thethrottle actuation arm 618, respectively. The servo-link 612 ispreferably manufactured from a length of steel wire, which is bent intoan aperture 620 formed through the servo-link arm 616 and secured forpivotal movement.

The throttle actuation arm 618 is positioned higher than the servo-linkarm 616, to provide clearance from the servo-link 612 when the enginethrottle 604 is actuated towards an open position. A slot 622 is formedthrough the throttle actuation arm 618, to allow the throttle arm 602 totravel in a relatively straight line of motion as the throttle actuationarm 618 rotates about the throttle link bolt 608. The slot 622 is openat the distal end of the actuation arm 618, to allow the throttle arm602 to be easily removed. The slot 622 also allows the engine 500 to beremoved from the vehicle without disrupting the throttle link assembly600, which is secured to the engine mount 510, rather than to the engine500.

The throttle 604 is actuated to an open position by servo-link 612pushing against the servo-link arm 616, rotating the bell crank 610 tomove the throttle actuation arm 618 towards the servo-link 612. Theservo-link 612 is secured by a guide 624 and stop 625 to a servoactuation arm 626 of a servo mechanism 613. The guide 624 allows theservo-link 612 to slide, while the stop 625 clamps the servo-link 612,preventing further sliding nearer the throttle 604.

The servo mechanism 613 rotates the servo actuation arm 626 about aservo mounting aperture 628 to move the actuation arm 626 towards thebell crank 610. The servo actuation arm 626 slides along the servo-link612 until the guide 624 abuts the stop 625, at which point, continuedmovement of the actuation arm 626 pushes the servo-link 612 to actuatethe bell crank 610. As the bell crank 610 actuates, the throttleactuation arm 618 moves towards the servo-link 612 and the throttle arm602 follows, opening the throttle 604. The guide 624 allows the servoactuation arm 626 to be actuated in an opposite direction, such as toactuate a braking mechanism (not shown), while leaving the throttle 604and the throttle link assembly 600 in the engine idle position (closed)shown. A spring 615 connected between an enclosure 617 holding the servoand the end of the servo-link 612 extending out of aperture 620 of thebell crank 610 returns the throttle 604 and a throttle link assembly 600to the engine idle position.

The configuration and position of the throttle link assembly 600 and theservo actuation arm 626 allow adjustment of the position of middlesupport 526 of the engine mount 510 and the engine 500, withoutrequiring decoupling of the throttle link assembly 600 from the engineor the servo actuation arm 626. Contributing to this is that the pivotpoints of the bell crank 610 and servo actuation arm 626 (excepting thepivot point at the throttle arm 602) form a substantially rectangularconfiguration in the unactuated position shown in FIG. 6D. Whenactuated, the pivot points form a trapezoid. In addition, the axis ofthe servo-link 612 is substantially perpendicular to the axis ofrotation of the bell crank 610 about the bolt 608. Thus, adjusting theposition of the engine 500 by the engine mount 510 does not requireadjustment of the throttle control link assembly 600.

FIGS. 9A through E illustrate a bumper assembly 650 that cooperates witha skid plate 652 to protect the front end of the vehicle shown fromimpacts. It will be apparent that the bumper assembly 650 may also bemounted on the rear end of the vehicle, to protect the back of thevehicle from impacts as well. The bumper assembly 650 comprises a bumpersupport 654 and a bumper 656 that are secured to a bumper chassis mount658 attached to the vehicle chassis 300. Below the bumper assembly 650and mounted to the bulkhead assembly 658 is the skid plate 652.

Referring additionally to FIG. 9E, the bumper support 654 is formed in agenerally oval-shape loop and is mounted to the bulkhead assembly 658 ina horizontal orientation relative to the chassis 300. The inboard length670 of the bumper support 654 includes two integrally formed mountingcollars 672 extending vertically across the width of the bumper support654. The mounting collars 672 are longer than the width of the bumpersupport 654, to provide greater resistance to and strength duringvertical flexing and twisting of the bumper support 654. The mountingcollars 672 extend vertically, to avoid interference with flexing of theinboard length 670 of the bumper support 654. A pair of fastening bolts673 extending through the mounting collars 672 and portions of thebulkhead assembly 658 secure the bumper support 654 to the front of thevehicle. The bumper support 654 also includes C-shaped, curved lateralends 674, each of which act as a curved leaf spring. The mountingcollars 672 are positioned to allow inboard deflection of the lateralends 674. The outboard length 676 of the bumper support 654 extendsbetween the lateral ends 674 and bends in a slightly convex curverelative to the bumper 656. The inboard and outboard lengths 670, 676 ofthe bumper support 654 also act as leaf springs to absorb an impact. Theoutboard length 676 of the bumper support 654 includes two integrallyformed mounting collars 678 extending horizontally and outwardly fromthe front of the bumper support 654. The mounting collars 678 preferablyextend outwardly from the outboard length 676 of the bumper support 654a sufficient distance to maintain clearance between the surfaces of thebumper 656 and the bumper support 654 in extreme impact conditions, whenmaximum deflection of the components occurs. The bumper support 654 ispreferably manufactured from a strong, elastic plastic, such as supertough Nylon® (Zytel ST 801), available from DuPont.

The bumper 656 is secured to the mounting collars 678 by a pair offastening bolts 680. The bumper 656 includes a frame member 682,surrounding a middle section of the length of the bumper 656. The framemember 682 adds rigidity and strength to the middle section of thebumper 656, as well as supporting a pair of substantially parallel,horizontally extending bumper stays 684. The outboard lengths of thebumper stays 684 each act as leaf springs to absorb an impact. Thebumper 656 is formed in a generally convex curve facing the front of thevehicle, to aid in deflecting the vehicle away from objects upon impactand to aid in deflecting movable objects from the path of the vehicle.The rear bumper can be flat, which is more stable for wheelies. Thebumper 656 is preferably manufactured from a strong, elastic plastic,such as super tough Nylon® (Zytel ST 801), available from DuPont.

The skid plate 652 is generally rectangular in shape, is substantiallyuniform in thickness and is secured to and extends forwardly from thebulkhead assembly 658. The skid plate 652 is positioned below andrearward of the bumper 656, and extends upwardly from the bulkheadassembly 658 toward the lower edge of the bumper 656. This orientationcauses the front surface of the skid plate 652 to face forwardly anddownwardly, to deflect obstacles away from the vehicle and to lift thevehicle's front end upwardly over obstacles in the path of travel. Theskid plate 652 acts as a leaf spring to absorb and protect the front endand bulkhead assembly 658 from impacts. Sufficient clearance is providedbetween the upper edge of skid plate 652 and the bumper 656, to avoidinterference as the skid plate 652 flexes. The skid plate 652 ispreferably manufactured from a strong, elastic plastic, such as supertough Nylon® (Zytel ST 801), available from DuPont.

In use, the bumper assembly 650 is capable of extreme deflection uponimpact. The outboard length 676 of the bumper support 654 will deflectinto contact with the inboard length 670, if necessary, on impact. Thelateral ends 674 will deform into a smaller radius, upon impact, whileboth the inboard and outboard lengths 670, 676 will deform or bowinwardly toward the center of the bumper support 654. Deflection of theoutboard length 676 of the bumper support 654 allows total deflection ofthe bumper support 654 in inboard direction greater than the deflectionof the lateral ends 674. The bumper support 654 will elastically returnto substantially the same position and shape following impact. The stays684 of the bumper 656 will also elastically deflect rearwardly, into amore bowed shape, upon impact. Following impact, bumper stays 684 willsubstantially return to the original shape.

Turning now to FIGS. 10-15, and initially to FIG. 10 thereof, aperspective view of a vehicle chassis 300 with the body shell 850removed is depicted, from the right side of the vehicle chassis 300.Vehicle chassis 300 has a fuel tank 852 secured thereon. Fuel tank 852has a fill opening 854 and a hinged filler cap 856. In one embodiment,the fill opening 854 has a rim 855 tipped toward a lateral side of thebody shell 850, at an angle with respect to the horizontal plane. In oneembodiment, this angle is between about 10 degrees and 80 degrees andmore preferably between about 40 degrees and 50 degrees. By making theopening 854 at an angle, the opening is more easily accessible for theoutside of the body shell 850 for filling. Furthermore, placing theopening 854 at an angle allows the fill opening 854 to be placed at theside of the body shell 850. The angle permits a fuel filler bottlenozzle to be inserted into the opening 854 without turning the bottleupside down over the vehicle, which reduces spillage. Furthermore, theangle makes the fuel cap easier to open by means of a direct upward pullon a finger ring pull, in a manner to be described below.

The angle also allows greater freedom of body shell styles since avertical opening would require a fuel neck extension to accommodatetaller body shell styles, such as SUV styles, or some other cumbersomemethod of refueling. However, with the angled opening, many body shellstyles of different heights can be used on the same chassis, withoutchanging the fill opening 854 or adding a fuel neck extension.

During fueling, air often becomes entrained in the fuel as it issqueezed into the tank, causing bubbles. These bubbles can cause foamand “burping” during filling, resulting in spills. To minimize thisproblem, the fuel tank 852 can include channels 853 along the insideupper surface of the top wall of the fuel tank 852, sloped upwardlyleading to the inside of the opening 854. These channels allow a pathfor entrained air in the tank to escape, toward the inside edges ofopening 854, where the escaping air is less likely to cause foaming or“burping” during filling.

The fuel tank 852 can have a resiliently closeable cap, such as a hingedfuel cap 856. Fuel cap 856 can be pivotably attached to molded eyes 857of the top of fuel tank 852 and attached with hinge pins 864. A spring866 can be installed between the fuel cap 856 and the tank 852 toresiliently urge fuel cap 856 into a closed position when it is notbeing intentionally physically opened for filling. The cap can also beclosed by a clip that snaps over the opposing end of the cap from thehinge and maintains the cap closed position.

Fuel cap 856 also includes a nozzle 858 to which is attached one end ofa pressurization tube 860. The other end of pressurization tube 860leads to a nozzle 861 on exhaust muffler 882. During operation of theengine 500, a slight amount of back pressure will be present in exhaustmuffler 882. Pressurization tube 860 causes this back pressure topressurize fuel tank 852, thus assisting fuel flow without the need torely on gravity alone and without the need for fuel pumps.

A finger pull tab 868 having an elongated shaft member 870 is attachedto the fuel cap 856. This pull tab 868 permits an operator to open thefuel cap 856 while keeping the users hands at a safe distance from hotor rotating objects that could injure them. This is advantageousbecause, after operation, the fuel cap can be soaked with fuel andsufficiently hot to risk injury from touching the fuel cap or, at theleast, an unpleasant burning sensation.

In accordance with an embodiment of the present invention, the fuel cap856 can be opened and closed, and the tank refilled, without the need toremove the body shell 850. However, if desired, the body shell 850 canbe removed and replaced for access to the fuel tank 852, or othercomponents on chassis 300, without the need to either open the cap 856or to remove the finger pull tab 868. However, as can be seen if FIG.12, the body shell 850 and the fill opening 876 in the body shell 850are spaced apart from opening 854 sufficiently so that the cap 856 canbe pulled open inside the shell 850 sufficiently to allow insertion of afuel filling line or nozzle, without removing the body shell 850. Asdepicted in FIG. 12, opening the cap 856 to an approximately horizontalposition is sufficient to provide substantially unimpeded access to theopening 854, but any degree of opening sufficient to allow insertion ofa fuel filling line or nozzle will suffice.

As can be seen in FIG. 12, the cap 856 can be opened by means of pullingup on finger pull tab 868, which extends through an opening 874 in thebody shell 850. Because FIG. 12 is a sectional view, only one half ofopening 874 is depicted, but it is to be understood that the remainderof the slot (not shown) is substantially a mirror image of the one halfof a opening 874 shown. Opening 874 is sized to permit the tab portion872 of pull tab 868 to pass without undue interference, to permitremoval and replacement of the body shell 850 without removal of pulltab 868. However, since pull tab 868 can be made from a resilientmaterial, such as plastic or rubber, some deformation of tab portion 872as it passes through opening 874 is permissible. Furthermore, having aseparate opening for the finger pull tab 868 provides greater access tothe fuel tank opening 854, since the finger pull tab 868 is safelyinside the slot 876, away from opening 854, and thus does not interferewith the fuel tank opening 854. The body shell 850 has a fill opening876 approximately aligned with the opening 854 in the tank 852.

Turning to FIGS. 16-18 and 19A-B, a vehicle chassis 300 having a secureddouble looped fuel line 800 in accordance with an embodiment of thepresent invention is depicted. Fuel line 800 has an intake end 802attached to a nozzle 804 which extends into fuel tank 852, from whichfuel can be withdrawn. Fuel line 800 has an exit end 806 that isattached to a carburetor 898 on engine 500. Fuel line 800 can be madefrom any suitable material, including a plastic or rubber materialgenerally resistant to the type of fuel employed.

As can be seen in FIG. 19A and B, the middle of fuel line 800 does notrun straight between the fuel tank 852 and the carburetor 898, butrather is coiled into a loop portion 808. In the event the vehicle turnsover during operation, fuel generally can no longer be drawn into theentrance of the fuel line 800. Accordingly, the engine will soon stoprunning. Normally, the vehicle will be operated by radio control and theoperator may be several hundred feet away from the vehicle at the timethe vehicle turns over. Often, this is too far to reach the vehicle toturn it upright before the engine stops. In the present invention, theloop portion 808 of the fuel line will retain additional fuel, givingthe operator additional time to reach and right the vehicle before theengine stops running from lack of fuel. It should be understood that,although a double loop is depicted, a single loop or more loops couldalso be employed.

Although the loop portion 808 will retain additional fuel, the coilingof the fuel line undesirably causes the fuel line to attempt to uncoil.Because the fuel line is nearby many hot surfaces, including the engine500 and exhaust pipe, the fuel line could easily come in contact withthese hot surfaces during rough drives. Accordingly, in accordance withthe present invention, the double loop is secured to the chassis byupper double clip 810 and lower double clip 812, which are affixed to asupport member such as roll bar 899 which is attached to chassis 300.

With the loop portion 808 secured, the advantages of using the loopportion 808 to provide additional fuel capacity in the fuel line isachieved, without the risk of fuel fires caused by unintended contactbetween the fuel line and a hot surface.

As can be seen in FIG. 17, the upper double clip 810 can have a firstfastener having a pair of opposed arcuate surfaces to grip a first loopof the loop portion 808 and a second fastener having a pair of opposedarcuate surfaces to grip a second loop of the loop portion 808. Thelower double clip 812 can have a third fastener having a pair of opposedarcuate surfaces to grip a lower portion of the first loop of the loopportion 808 and a fourth fastener having a pair of opposed arcuatesurfaces to grip a lower portion of the second loop of the loop portion808. At least a portion of one of the opposing surfaces of the thirdfastener is spaced farther from the other opposing surface to receiveand retain the curved surface of a portion of the tube retained by thethird fastener. Also, at least a portion of one of the opposing surfacesof the fourth fastener can be spaced farther from the other opposingsurface to receive and retain the curved surface of a portion of thetube retained by the fourth fastener.

The first and third fasteners can be formed as one integral piece andthe second and fourth fasteners can also be formed as one integralpiece. Thus, the third fastener can form an entrance for placement of aportion of a tube in the first fastener and the fourth fastener can format least a portion of an entrance for placement of a portion of a tubein the second fastener. Conveniently, either or both double clips 810and 812 can be molded integrally with roll bar 899, which isconveniently made of a plastic material. Because both the fuel line 800and the double clips 810 and 812 are somewhat resilient, the fuel linescan be resiliently inserted into the clips and resiliently retainedthere during rough driving, while still being removable intentionally bythe operator without difficulty

FIGS. 20A-C through 24 illustrate a slipper clutch assembly 900 fortransferring torque from the spur gear 518 shown in FIG. 1 to atransmission input shaft 902, during operation of the vehicle. Theslipper clutch assembly 900 protects the spur gear 518 and the engine500 shown in FIG. 1 from acute shocks to the drive train, such as whenthe wheels of the vehicle are abruptly slowed from a high speed spin toa much lower rotation when the vehicle lands following a jump. Theslipper clutch can also serve as a torque limiting traction control aid.The slipper clutch assembly 900 interposes a friction coupling betweenthe spur gear 518 and the transmission input shaft 902, whichmomentarily slips, allowing the spur gear 518 to rotate at a speedfaster than the input shaft 902 until the speed is slowed by thefriction coupling of the slipper clutch assembly 900. When acute shocksto the drive train are not experienced, the slipper clutch assembly 900preferably transmits rotational torque with little or no slippage.

The slipper clutch assembly 900 is configured to allow removal of thespur gear 518 without changing the compression setting of the slipperclutch assembly 900. The spur gear 518 is secured directly to the driveplate 904 by bolts 906 extending through substantially equidistantlocations on the body of the spur gear 518. The bolts 906 are threadedinto similarly located receptacles 908 formed on the surface of thedrive plate 904. The spur gear 518 can be removed from the slipperclutch assembly 900, for service or replacement, by removing the bolts906 from the receptacles 908.

The slipper clutch assembly 900 transfers torque between the spur gear518 and the input shaft 902, depending upon the compressive forceapplied to the drive plate 904 and the driven plate 910. The compressiveforce is adjusted by an adjustment nut 912 threaded on the end of theinput shaft 902 extending from the vehicle transmission (not shown). Theadjustment nut 912 abuts and compresses a pair of springs 916 mounted onthe input shaft 902 to maintain the desired compressive force. Althoughsprings 916 are spring washers, it will be apparent that other suitablesprings, such as helical springs and the like, could be employed. Thesprings 916, in turn, press a radial ball bearing assembly 918 againstthe drive plate 904. The drive plate 904, in turn, presses clutch pads920 against a clutch disc 922 held by the driven plate 910 of theslipper clutch assembly 900. Frictional resistance to movement betweenthe contacting surfaces of the clutch pads 920 and the clutch disc 922couples the spur gear 518 to the transmission input shaft 902. Therotational and axial position of the driven plate 910 is secured by apin 926 that extends through a diametrically extending hole 928 throughthe transmission input shaft 902. Opposing ends of the pin 926 extendfrom the hole 928, against the driven plate 910 and prevent movement ofthe plate axially along the shaft 902 away from the adjustment nut 912.The greater the compressive force applied to the clutch pads 920 and theclutch disc 922, the more torque will be required to cause slippage ofthe slipper clutch assembly 900.

The ball bearing assembly 918 supports the spur gear 518 for rotationabout the transmission input shaft 902, in addition to transmittingcompressive forces from the spring(s) 916. An aperture 924 in the centerof the spur gear 518 preferably fits snugly over the ball bearingassembly 918. The ball bearing assembly 918 also fits snugly over thetransmission input shaft 902. This configuration reduces the totalclearance encountered between the input shaft 902 and the teeth of thespur gear 518, reducing the risk of run out by the spur gear 518.

The clutch pads 920 are each supported by a flange 929 extendingoutwardly from a central, circular body portion of the drive plate 904.The clutch pads 920 each include a pair of indexing holes 930 in theirsurfaces opposite the clutch plate 922. Indexing posts 932 extendingfrom the flanges 929 insert into the indexing holes 930, secure theclutch pads 920 from sliding out of position during operation.

The clutch disc 922 is secured against movement by the driven plate 910of the slipper clutch assembly 900. The clutch disc 922 has a circularouter perimeter substantially matching the circular perimeter of thedriven plate 910. However, a central portion is cut from the clutch disc922 in an irregular pattern, substantially matching a similar pattern934 extending from the surface of the driven plate 910 toward the driveplate 904. The perimeter of the irregular pattern cut in the clutch disc922 fits around the similar pattern extending from the driven plate 910,to secure the clutch disc 922 for rotation with the driven plate 910.

The driven plate 910 is secured for rotation with the transmission inputshaft 902 by the pin 926, the ends of which engage an opposing pair ofslots 936 formed in a collar 938 extending around the input shaft 902and away from the drive plate 904. The pin 926 and the slots 936cooperate to index rotation of the driven plate 910 to the input shaft902. Rotation of the driven plate 910 rotates both the pin 926 and theinput shaft 902.

Extending from the surface of the driven plate 910 are a number ofintegrally formed vanes 940. The vanes 940 trace spiral paths outwardlyover the area of the driven plate 910 supporting the clutch disc 922. Asthe driven plate 910 rotates, the spiral vanes 940 act as cooling finsto dissipate heat caused by friction between the clutch disc 922 and theclutch pads 920 during operation of the vehicle.

The slipper clutch assembly 900 provides reduced size, low inertia andenhanced heat dissipation. These features are provided by use of asemi-metallic, high-friction material to form the clutch pads 920. Useof such a high-friction material allows placement of the clutch pads 920closer to the axis of rotation of slipper clutch assembly 900, reducingthe diameter of the slipper clutch assembly 900. The reduced diametercontributes to both reduced size and low inertia. Both the drive anddriven plates 904, 910 are preferably manufactured from cast aluminum,which is light-weight and a good heat conductor, further contributing tolow inertia and enhanced heat dissipation.

In prior model vehicle braking pad assemblies, a thin piece of frictionmaterial is supported by a pad support constructed of a thin piece ofsheet metal. A small piston, actuated by a cam, applies force to thesheet metal plate. The plate applies force to the friction material anddisk. A problem with such prior braking pad assemblies is that the useof thin and flexible material for the pad support and friction materialresults in poor distribution of pressure, overheating and uneven wear.As a result, the area directly under the piston wears quickly andoverheats.

In order to overcome these disadvantages of prior model vehicle brakingpad assemblies, in an embodiment of the present invention, the frictionmaterial can be supported by a very rigid cast pad holder (also called acaliper). The pad holder geometry is more three dimensional than typicalpads that are stamped from sheet metal. This structure also provides thecaliper with a high thermal capacity and better thermal conductivity forcooling. Furthermore, in an embodiment of the present invention, thecaliper can employ an integrated post with ribs providing additionalstiffness to help evenly distribute the forces from the actuating cam.In another embodiment, an integrated cam receiving surface on thecaliper also helps to evenly distribute the forces from the cam.

FIGS. 25A-D, 26A-B and 27-28 depict a model vehicle braking pad caliberassembly 1000 in accordance with in an embodiment of the presentinvention. The braking pad caliper assembly 1000 has outboard pad madeof a friction material 1002 supported by a very rigid cast pad holder orcaliper 1004 on the outboard side of braking disk 1006. On the inboardside, an embodiment of the invention can include a pad of frictionmaterial 1008 supported by an opposing very rigid cast pad holder orcaliper 1010 on the inboard side of braking disk 1006. The braking disk1006 can be made from strong material, such as steel, aluminum ortitanium. The braking disk further can have slots 1001 and holes 1003for, respectively, reduction of weight and assisting cooling of thedisk. The calipers 1008 and 1010 can be made from a strong material,such as steel, aluminum or titanium. In an embodiment, the calipers 1008and 1010 can be made from cast aluminum, which has a higher thermalconductivity than steel as well as a high strength to weight ratio.

Disk 1006 is slidably mounted over drive shaft 1012 but not affixed toit. That is, the disk 1006 is free to slide axially on the shaft 1012 toa limited degree. Drive shaft 1012 has opposite flat surfaces 1013 and1015 on its end 1011 for receiving a coupling (not shown). The couplinghas two pin keys (not shown) that extend into opposite ends 1018 and1020 of slot 1022, that extends from hole 1017 in disk 1006. These pinkeys force the disk 1006 to rotate with the coupling, and hence with thedrive shaft 1012 but permit a limited degree of axial sliding of thedisk 1006 with respect to drive shaft 1012.

As can be seen in FIG. 27 and FIG. 28, in one embodiment, the brake padsupport calipers 1004 and 1010 each support a brake pad of frictionmaterial 1002 and 1008 on first inner faces 1005 and 1009, respectively,to which the friction material 1002 and 1008 is disposed. In oneembodiment, the calipers 1004 and 1010 can each be a single piece ofcast aluminum.

In one embodiment, the inboard caliper 1010 has a cam receiving post orfollower 1016 extending from its outside face 1045. The post 1016 has acam receiving surface for receiving compressive force from an actuatingcam 1025.

The actuating cam 1025 can take a variety of forms. In one embodiment,the cam 1025 is the flat surface 1027 of a half-shaft portion of a camshaft 1023. The cam shaft 1023 is retained in base 1032 for pivotingabout the axis of cam shaft 1023. In one embodiment, base 1032 is thetransmission housing, which is secured to chassis 300. The cam shaft ispivoted by means of a force applied to yoke 1021, which is secured toone of the ends of cam shaft 1023.

As the cam shaft is pivoted, one side of the flat surface 1027 willcompressively press against the cam receiving surface of post 1016. Thiswill, in turn, displace the inboard caliper 1010 and the frictionmaterial 1008 on it toward the disk 1006.

The brake calipers 1004 and 1010 can further include a plurality offastening points 1024, 1026 and 1028 and 1030 at which the respectivecaliper is secured directly or indirectly to the chassis 300 of a modelvehicle. As can be seen in FIGS. 25A-D, for example, the fasteningpoints 1024 and 1026 for the outboard caliper 1004 are where the caliperis attached to the base 1032 by means of screws through screw holes 1051and 1053. In the case of the inboard caliper 1010, the caliper hassecuring holes 1028 and 1020 at each of its ends, which can slide overthe shafts 1038 and 1040 of securing screws 1034 and 1036. However, thecaliper 1020 is not fixedly secured to the shaft portion of the screws,but instead is axially free to slide along the shafts of the screws sothat the friction material disposed on the caliper can be pressedagainst the disk 1006 during brake actuation.

As indicated above, the disk 1006 is free to slide axially to somedegree along the axis of drive shaft 1012. Thus, as the inboard caliper1010 and its friction material 1008 are forced toward the disk 1006, thedisk will be free to slide towards the friction material 1002 on theoutboard caliper 1004, which is fixed in place by means of the heads ofthe screws 1034 and 1026 securing it to base 1032. Thus, when the brakeis actuated by the cam, the axially slidable disk 1006 will be“sandwiched” in between the movable inboard caliper 1010 and the fixedoutboard caliper 1004, effectively applying braking force to stoprotation of the disk. This will stop rotation of the drive shaft 1012which will also cause stopping of the rotation of all the wheels (notshown) connected to the drive shaft.

As can be seen in FIGS. 25 A through D, 26A and B, 27 and FIG. 28, oneor more ribs 1042 and 1007 extend outwardly across substantially theentire length of the outer surface of the caliper 1004. The term“inner”, when referring to either caliper 1004, 1010, means the surfacein contact with the friction material. “Outer” means the other surfaceof the caliper plate 1004, 1010. Ribs 1007 extend substantially parallelto the circumference of an axle of shaft 1012 to be braked, while ribs1042 extend substantially tangentially to the circumference of the axleor shaft 1012. The ribs 1042 act to stiffen the caliper 1004 todistribute compressive forces applied to the outside face at one or morelocations on the caliper, as well as to provide cooling. As can be seenbest in FIG. 25C, one or more of the ribs 1042 can be tapered in heightas the rib approaches one of the plurality of plate fastening points1034, 1036. Thus, the ribs 1042 are the highest at the middle of thespan, where the bending moment would be the highest. Furthermore, theone or more ribs 1042 extend across at least a portion of the outerfaces of the calipers in substantial alignment with an imaginary linedrawn through the center point of each of the plurality of fasteningpoints 1034 and 1036. The plurality of ribs 1007 extend across at leasta portion of the outer surface of the calipers 1004 and 1010, which canfacilitate cooling of the calipers, as well as providing stiffeningreinforcement. The ribs 1007 can each extend from the nearest rib 1042on the outer surface of caliper 1004 to curve circumferentially aboutthe axis of drive shaft 1012 toward an edge of caliper 1004, thusproviding additional stiffness in the direction of applied frictionalforce, in addition to providing cooling.

In order to retain the friction material 1002 and 1008 in position onthe respective calipers, the calipers can include one or more brake padbosses 1048 extending from the inner face of the caliper for engaging atleast a portion of the perimeter of a pad of friction material 1002 or1008 supported on the inner face of the caliper, to resist lateralmovement of a brake pad 1002 or 1008 across the inner surface of therespective caliper. The bosses 1048 have space between them so that anoperator can visually determine the degree of wear of friction materialwithout the need for disassembly. The brake pad bosses 1048 can besufficient alone to retain the friction material in position on thecaliper without the need for reliance on other means for fastening thefriction material to the caliper. However, if desired, the frictionmaterial can also be secured to the caliper by adhesive, screws, rivetsor other convenient means

Co-pending U.S. Patent Application of Brent W. Byers entitled “A ModelVehicle Suspension Control Link” (Docket No. TRAX 3175000), filedconcurrently herewith, is hereby incorporated by reference for allpurposes. Components depicted in this application having substantiallysimilar construction and function to those shown in the co-pendingapplication hereby incorporated by reference are identified with thesame reference numeral, followed by a prime (′) designation (e.g.,100′). For example, various components employed in the construction andoperation of the rear suspension arm assembly 100 in the co-pendingapplication are substantially similar in construction and operation tothe components employed in the front suspension arm assembly 100′ shownin FIGS. 29A through D.

Referring now to FIGS. 29A through D, shown is a front bulkhead assembly658, from which laterally extends a suspension arm assembly 100′ and atelescoping drive shaft 1100. The telescoping drive shaft 1100 extendsand retracts with upward and downward movement of the suspension armassembly 100′. The drive shaft 1100 is secured by a Cardan joint 1102(sometimes referred to as a “universal joint”) to a transmissiondifferential assembly shown in FIGS. 29A-D mounted in a fixed positionon the front bulkhead assembly 658. The outboard end of the drive shaft1100 is secured by a Cardan joint 1102 to an axle assembly 1104 (shownin one or more of FIGS. 33D, 34 and 35) mounted for rotation within anaxle carrier 140′. The axle carrier 140′ is supported on the outboardend of the suspension arm assembly 100′. Extension and retraction of thetelescoping drive shaft 1100 accommodates a different pivotal pathfollowed by the axle carrier 140′ as the suspension arm assembly 100′moves between uppermost and lowermost positions.

Referring now to FIGS. 30A through D, 31A and B, and 32A and B, thetelescoping drive shaft 1100 is shown in greater detail. The drive shaft1100 comprises an inboard yoke 1106 for securing a tubular externalsegment 1108 to the front transmission differential of the vehicle. Anoutboard yoke 1110 forms the outboard end of the drive shaft 1100 forsecuring a tubular internal segment 1112 to the Cardan joint 1102coupling of the drive shaft 1100 to the axle assembly 1104. The inboardand outboard yokes 1106, 1110 are integrally formed with the remainderof the external and internal segments 1108, 1112, respectively, in asingle-piece construction.

As is best shown in FIGS. 32A and 32B, curved splines 1114, 1116 extendfrom the internal and external surfaces, respectively, of the externalsegment 1108 and the internal segment 1112 of the drive shaft 1100. Thesplines 1114, 1116 extend at least along the lengths of the external andinternal segments 1108, 1112 that will overlap when the suspension armassembly 100′ travels between the uppermost and lowermost positions. Thesplines 1114, 1116 are aligned with the longitudinal axis of the shaftsegments 1108, 1112, respectively, in a parallel formation. In theembodiment shown, the splines 1114 extend along substantially the entirelength of the inner wall of the external segment 1108. The curvedsurfaces of the splines 1114, 1116 are complementary, each mating with acorresponding groove formed between adjacent splines of the external andinternal segments 1108, 1112, respectively. The splines 1114, 1116 varyin radius of curvature at approximately 180° intervals about therotational axis of the drive shaft 1100. In the embodiment shown, forexample, indexing splines 1118 of the external segment 1108 and indexingsplines 1120 of the internal segment 1112 have a smaller radius ofcurvature relative to other of the splines 1114, 1116. The radius ofcurvature of the corresponding grooves with which the indexing splines1118, 1120 mate, have a similarly smaller radius of curvature. Thisindexes the external and internal segments 1108, 1112 when mated, toassure alignment of the yokes 1106, 1110 in substantially the samerotational position.

The curved splines 1114, 1116 transfer torque between the yokes 1106,1110, while allowing the segments 1108, 1112 of the drive shaft 1100 toslide with respect to each other, in telescopic fashion. The curvedsurfaces of the splines 1114, 1116 allow more splines to be formed thanif rectangular splines were used. The curved surfaces and number of thesplines 1114, 1116 and corresponding grooves reduce or eliminate stressconcentrations experienced by telescopic drive shafts employingrectangular splines. Stress reduction and accommodation of a greaternumber of splines 1114, 1116 is provided by a relatively larger thantypical diameter employed by the drive shaft 1100. These attributes alsoallow the walls of the internal and external segments 1108, 1112 to bethinner and lighter in weight.

The segments 1108, 1112 of the drive shaft 1100 are preferablymanufactured from a low-friction, high impact strength plastic, or othersimilar material. In the embodiment shown, the segments 1108, 1112 aremade from a suitable Nylon material. The low-friction attributes ofthese materials substantially eliminates the need to lubricate thesurfaces of the segments 1108, 1112.

The drive shaft 1100 is sealed to prevent dust, dirt, debris and thelike from entering and causing abrasion of and friction between thesurfaces of the segments 1108, 1112, which would reduce performance andlongevity. The ends of the drive shaft 1100 next to the yokes 1106, 1110each include respective apertures 1122, 1124 that are sealed byelastomeric plugs 1126, 1128 secured by a compression fit. The seambetween the surfaces of the external and internal segments 1108, 1112 issealed by a bellows seal 1130.

The bellows seal 1130 includes a substantially cylindrical centralportion 1132, having laterally extending folds, allowing both expansionand retraction of the bellows seal 1130 with expansion and contractionof the drive shaft 1100. Extending from the inboard and outboard ends,respectively, of the bellows seal 1130 are substantially cylindrical,smooth sealing collars 1134, 1136. The sealing collars 1134, 1136,respectively, fit snugly over substantially cylindrical, smooth landingsurfaces 1138, 1140 formed on the external surfaces of the segments1108, 1112. A seal is formed between the sealing collars 1134, 1136 andthe landing surfaces 1138, 1140, by a compression seal. In addition, thesealing collars 1134, 1136 are secured to the landing surfaces 1138,1140, by a suitable glue or adhesive. The bellows seal 1130 ispreferably made from a suitable rubber compound, such as nitrile rubber,and the like.

FIGS. 33A through D, 34 and 35 illustrate coupling of the drive shaft1100 via the Cardan joint 1102 to a drive axle assembly 1104 for drivinga wheel 120′ on the front end of the vehicle. The Cardan joint 1102comprises the outboard yoke 1110 of the drive shaft 1100 coupled to adrive axle yoke 1142. The drive axle assembly 1104 is supported by theaxle carrier assembly 140′ for rotation. A drive pin 1144 couples thedrive axle yoke 1142 to the drive axle assembly 1104 to transfer torquefrom the drive shaft 1100 to the wheel 120′. The drive axle yoke 1142 issupported for rotation within the axle carrier 140′ by an internallymounted radial ball bearing assembly 1146. Supporting the drive axleassembly 1104 for rotation is a ball bearing assembly 1148 mounted inthe axle carrier 140′ adjacent the wheel 120′.

In addition to transferring torque from the yoke 1142 to the axleassembly 1104, the drive pin 1144 secures the yoke 1142 to the axleassembly 1104. The drive pin 1144 comprises a substantially smooth,cylindrical pin extending through an aperture extending diametricallythrough the outboard shank of the drive axle yoke 1142 and an alignedaperture extending diametrically through a portion of the axle assembly1104 inserted into the shank. The interior surfaces of the apertures ofthe shank of the drive axle yoke 1142 and the axle assembly 1104 arepreferably smooth and provide sufficient clearance to allow the drivepin 1144 to be inserted and removed without difficulty.

The ball bearing assembly 1146 serves the dual purpose of supporting thedrive axle yoke 1142 shank for rotation and securing the drive pin 1144within the shank. This configuration allows replacement of the driveaxle yoke 1142, for example, if damaged, without the need to replace thedrive axle assembly 1104 as well. Various manufacturing steps andassociated costs are also reduced or eliminated

FIG. 36 illustrates substantially identical ball joint assemblies 1150pivotally supporting the axle carrier 140′ on the outboard ends of theupper and lower suspension arms 102′, 104′. In FIGS. 36 and 37, the yoke1142, axle assembly 1104 and related components have been removed. Theball joint assemblies 1150 allow universal movement of the axle carrier140′ relative to the suspension arms 102′, 104′ to allow steering, wheelalignment and suspension travel.

The ball joint assemblies 1150 each include a substantially sphericalball 1152 having a threaded shank 1154 securing each of the balls 1152to one of the suspension arms 102′, 104′. Formed into each of the balls1152 is a socket 1156, preferably hexagonal, substantially aligned withthe central axis of the threaded shank 1154. The socket is used tosecure the shanks 1154 to the suspension arms 102′, 104′ and to adjustthe distance between the balls 1152 and the suspension arms 102′, 104′.Adjustment of the balls 1152, in turn, allows adjustment of the camberof a wheel supported by the suspension arms 102′, 104′, in particular.Removal of the balls 1152 from their respective suspension arms 102′,104′ facilitates maintenance and replacement of parts.

An inboard portion of each of the balls 1152 slides into acorrespondingly shaped inboard end of a ball housing 1158. Each ballhousing 1158 is generally cylindrical and extends from the outboardsurface of the axle carrier 140′, beginning with a diameter large enoughto accommodate insertion of the ball 1152 and forming a substantially aspherical surface ending in an inboard aperture through which the ballshank 1154 extends. Formed in the surfaces of each housing 1158 arethreads 1160 for receiving and securing a pivot ball cap 1162 forretaining each ball 1152 within the respective housing 1158.

Each pivot ball cap 1162 is generally tubular, having external threads1164 mating with housing threads 1160 and an inboard bearing surface1166 for securing a ball 1152 within the respective housing 1158. Thebearing surface 1166 is formed about the open, inboard end of each cap1162 and is substantially flush with the spherical surface of theassociated ball 1152. The pivot ball caps 1162 are tightened to justtake up excess clearance with the balls 1152, the threads have a mildinterference fit with the housing threads 1160 to prevent loosening ofthe caps 1162. Removal of the caps 1162 allows the balls 1152 to beremoved from the housings 1158 for maintenance, repair and replacement.Extending from the perimeter of the outboard end of each of the caps1162 are a number of fingers 1167, forming a castle gear that is used tothread and unthread each of the caps 1162. It will be apparent that thenumber of fingers 1167 and their configuration may be varied, asdesired.

Seated in each cap 1162 is a self-healing cap seal 1168 to prevent dust,debris, dirt and other contaminants from entering the housings 1158.Each cap seal 1168 includes a head portion 1170 having a radial lipextending to the fingers 1167 of the cap 1162. The head portions 1170rest on and form a seal against the throat portions 1172 of the caps1162 extending inwardly and inboard of the fingers 1167, forming alanding for the head portions 1170. Extending from the head portion 1170of each cap seal 1168 is a neck 1174 extending through and contactingthe surfaces of the cap throat portion 1172, forming a further seal.Each cap seal 1168 includes a retaining lip 1176 extending radially fromthe neck 1174 to assist in retaining the seal within the respective cap1162. The cap seals 1168 are preferably manufactured from a pliablenitrile rubber that can be deformed, but will elastically return to theoriginal shape.

Formed in the head portion 1170 of each cap seal 1168 is a self-healingaperture 1178. The self-healing aperture 1178 is preferably formed by apair of slits cut through the head portion 1170 intersecting atsubstantially 900. The slits normally abut to maintain a seal. However,a hexagonal wrench, lubricating nozzle or other tool can be insertedthrough the self-healing aperture 1178, parting the lips of the slits,to adjust, remove, maintain or lubricate the associated ball 1152. Whenthe tool is removed, the self-healing aperture 1178 elastically returnsto the original, sealed position.

The inboard end of each housing 1158 is sealed by an elastic boot 1180that extends between the shank 1154 of each ball 1152 and a landing 1182formed on the axle carrier 140′ about the inboard aperture of the balljoint housing 1158. Each boot 1180 is generally conical in shape,extending from a wider opening adjacent the axle carrier 140′, to asmaller opening that surrounds the associated shank 1154. Each boot ispreferably manufactured from a material similar to that of the cap seals1168. The walls of each boot preferably form a number of folds, allowingthe boot 1180 to flex easily with movement of the axle carrier 140′, andwithout tearing or binding.

Referring now to FIGS. 37, 38 and 39 A through C, each boot 1180 issecured to the landing 1182 by a ring 1184 which fits over andcompresses a cylindrical portion of the boot 1180 into sealingengagement with the landing 1182. A lip 1186 extends radially from thecylindrical portion of the boot 1180 and is compressed against ashoulder 1188 formed on the surface of the axle carrier 140′. Each ring1184 is held in this position by a pair of clips 1190 extendingsubstantially perpendicularly from and on diametrically opposed pointson the ring. The clips 1190 are pressed over a pair of clip receptacles1192 positioned on opposite sides of the associated ball housing 1158.The rings 1184 and clips 1190 are preferably manufactured from a strong,impact-resistant plastic.

The inboard ends of the boots 1180 are each secured to the associatedshanks 1150 by an elastic collar 1194 integrally formed at the narroweropening of each of the boots 1180. The elastic collars 1194 aresubstantially thicker than the walls of their respective boots 1180 andform a compression seal against the underlying surface of the associatedshank 1154. Each collar 1194 is retained by an annular insert 1196formed about the circumference of the associated shank 1154 at alocation preferably outboard of the respective suspension arms 102′,104′. The shoulders of the annular inserts 1196 retain the collars 1194from sliding over the associated shanks 1154

Turning now to FIGS. 40A-D, 41A-B and 42, a dual arm centrally mountedsteering arm 1200 driven by a pair of servos 1202 is depicted. Thecentrally mounted steering arm 1200 is pivotally mounted to amountingbracket 1204 by means of a mounting screw 1206, which passes through abushing 1208, a center hole 1207 in a retainer 1209, and a center hole1210 in steering arm 1200.

At each of the ends 1211 of steering arm 1200 are yokes 1212, to whichcan be attached a rod assembly 1214. Each rod assembly 1214 includes twoball joint ends 1216 and a center rod portion 1218. In one embodiment,the ball joint ends 1216 employ hollow ball bushings 1220. One of theball joint ends 1216 is pivotally connected to one of the yokes 1212 bymeans of screw 1222, which passes through the yoke 1212 and through thehole in the hollow ball bushing 1220. The other of the ball joint ends1216 is pivotally connected to an actuator arm 1217 of one of the pairof servos 1202 by means of screw 1219 through yoke 1225 at the end ofactuator arm 1217. Actuator arm 1217 is, in turn, attached to the outputshaft 1224 of the servo by means of attachment screw 1226.

In operation, when the operator desires to turn the vehicle, a signal issent to both of the servos 1202 at substantially the same time. Each ofthe servos 1202 will cause their output shafts 1224 to pivot in oppositedirections, at about the same time. This will cause rod assembly 1214 toextend and retract, applying force to the yokes 1212 of the steeringarm, respectively, pivoting the centrally mounted steering arm 1200.

In order to minimize the potential for damage to the servos or theirconnecting rods and arms, a spring and cam servo saver 1240 assembly isused to connect to a driven steering arm 1242. Driven steering arm 1242is, in turn, connected to a pair of hollow ball end steering control tierods 1244, one of which controls the steering position of one of the twofront wheels 120′, and the other of which controls the steering positionof the other of the two front wheels. The ball end of each of the tierods 1244 is attached to an end 1246 of driven steering arm 1242 bymeans of screws 1248. Driven steering arm 1242 pivots about bushing1208, which passes through a hole 1250 in driven steering arm 1242.

The servo saver assembly includes retainer 1209, spring 1252, centrallymounted steering arm 1200 and driven steering arm 1242. Centrallymounted steering arm 1200 includes a pair of axially rotable arcuatelugs 1254, which act as cam surfaces, which fit into cooperativelydesigned hollows 1256 in the facing surface of driven steering arm 1242,which act as mating cam surfaces. Retainer 1209 is then secured todriven steering arm 1242 by means of screws 1258, with conical spring1252 resiliently urging centrally mounted steering arm 1200 againstdriven steering arm 1242 so that lugs 1254 center themselves intohollows 1256.

Under normal steering, the resilient force of spring 1252 is sufficientto keep lugs 1254 in place in hollows 1256 so that pivoting of centrallymounted steering arm 1200 by driving it with servos 1202 will causedriven steering arm 1242 to simultaneously pivot, ultimately resultingin steering of the wheels through steering control links 1244. However,when the vehicle wheel strikes an obstruction during rough driving forexample, excessive forces can be imposed on the steering components thatmight cause damage to the components. When this occurs, the drivensteering arm 1242 will pivot relative to centrally mounted steering arm1200, causing lugs 1254 to rise out of the hollows 1256 against theresilient force of spring 1252. This relative pivoting limitstransmission of force from driven steering arm 1242 to the rest of thesteering components, thus minimizing the potential for damage. However,immediately upon removal of the excessive force, the lugs 1254 will“pop” back into hollows 1256 under the resilient force of spring 1252,thus returning the steering assembly to normal operation.

By use of a pair of servos 1202 mounted on the left and right side ofthe chassis 300, a symmetrical torque is applied to the steering arm1200. This results in a huge benefit to performance minded users due tocrisp break away, strong centering and less looseness and/or hysteresisin the system. Furthermore, use of a centrally mounted steering armpermits use of a single, central servo saver, instead of a separateservo saver for each servo, eliminating additional parts and loosenessand/or hysteresis in the system

Turning now to FIGS. 43A-D and 44-46, a mounting system for securelymounting a servo 1202 to the chassis 300 by means of a clamp stylebracket 1300 and a clamp style bracket 1301 is depicted. Servo 1202includes a housing 1302, which can conveniently be molded of plastic.Housing 1302 includes attachment ears 1304 extending from the endsthereof, which can conveniently be molded integrally with the ends ofhousing 1302.

Rather than attach the attachment ears 1304 directly to the chassis 300by means of screws, for example, as is conventional, in accordance withthe present invention, a clamp style forward bracket 1300 and a clampstyle aft bracket 1301 are employed to secure the attachment ears to thechassis 300. Forward bracket 1300 has an upper flange 1306 and a lowerflange 1308. Upper flange 1306 has a pair of threaded holes 1309 whichare adapted to receive the threaded end of a screw 1311. Upper flange1306 and lower flange 1308 are connected at one end by an arcuate livehinge 1310, which can conveniently be molded integrally with upperflange 1306 and lower flange 1308 from plastic material. In addition,lower flange 1308 can includes one or more downwardly extending bossportions 1329A and 1329B, which extend below the upper surface ofchassis 300, into the openings 1307A and 1307B of the chassis, to fixthe forward bracket 1300 against forward/aft movement. Lower flange 1308has a hole 1313 disposed through it for accepting the shaft 1315 ofscrew 1311. Hole 1313 need not be threaded.

Aft bracket 1301 has an upper flange 1316 and a lower flange 1318. Upperflange 1316 has a pair of threaded holes 1319 which are adapted toreceive the threaded end of a screw 1311. Upper flange 1316 threaded andlower flange 1318 are connected at each of their sides by an arcuatelive hinge 1320, which can conveniently be molded integrally with upperflange 1316 and lower flange 1318 from plastic material. Lower flange1318 can have one or more downwardly extending lateral bosses 1330 and1331, which extend below the upper surface of chassis 300, intorespective openings 1333 and 1335 of the chassis, to fix the aft bracket1300 against forward/aft movement. Lower flange 1318 has a hole 1323disposed through it for accepting the shaft 1325 of screw 1311. Hole1323 need not be threaded.

To secure the body 1302 of servo 1202, forward bracket 1300 is put ontothe end of one of the attachment ears 1304, and bracket 1301 is put ontothe end of the other of the attachment ears 1304. Then, screws 1311 aresecured, securely clamping one of the ears 1304 between upper flange1306 and lower flange 1308, and the other of the ears between upperflange 1316 and lower flange 1318.

Brackets 1300 and 1301 can be manufactured from Zytel 70 G 33 (33%Glass) available from DuPont, which retains shape and grips screwthreads better than plastics without a glass reinforcing fill.

By use of the clamping type brackets of an embodiment of the presentinvention, a wide range of aftermarket dimensions of servos can beaccommodated without requiring additional parts and without compromisein the mounting integrity. Furthermore, the clamp style interfacedistributes loads over the entire mounting ear thereby reducingbreakage/distortion of the mounting ears, overall improvement indurability. In addition, the clamp style mounting type brackets alsoimprove control performance by increasing the stiffness of theservo-vehicle interface. Of course, the forward and aft brackets couldbe reversed, if desired

FIGS. 47A and B illustrate a vehicle 1400 incorporating the variousfeatures described herein, including in Appendices A, B, C and D hereto,which are incorporated herein by reference.

Referring now also to FIGS. 1 and 47A through 52, illustrated is achassis 300, which is also described elsewhere in connection with otherfeatures and components comprising portions of the vehicle 1400. Thechassis 300 is configured to provide a lower center of gravity than cantypically be provided by conventional chasses resembling a relativelyflat surface or plate. This is accomplished by providing chassis 300with flanges 302 extending laterally from a central channel area 304.The lateral flanges 302 extend from downwardly sloping lateral walls 306of the central channel area 304 at a substantially lower level relativeto an underlying surface. The lateral flanges 302 provide support forrelatively heavy components that do not require placement near or inalignment with the drive train of the vehicle 1400. In general, theflanges 302 lower the mounting points of various components on thechassis 300, at least relative to the transmission assembly 520 andtransmission output shaft 521. In addition, the flanges 302 preferablyincline gradually as they extend laterally from the channel area 304.Upward sloping of the flanges 302 causes the components supported on theflanges 302 to extend both upwardly and inwardly toward the center ofthe vehicle 1400, more tightly packaging the components on the chassis300.

The flanges 302 preferably include openings 308, for example, throughwhich the lower portions of components can extend, in addition to beingsecured to the flanges 302 at a lower level than the central channelarea 304. Where convenient, chassis 300 weight is reduced by configuringone or more flanges 302 as a support arm, such as arms 302A, thatcooperates with other flanges 302 to support components on the chassis300. Further, the flanges 302 may preferably extend laterally andsubstantially without upward inclination, if desired to enhanceperformance of the component or to satisfy structural or packagingpreferences.

The flanges 302 are capable of supporting numerous components of thevehicle 1400 at a level substantially lower than the central channelarea 304. In the embodiment shown, the flanges 302 support at a lowerlevel, an electronics and battery package 1402, a fuel tank , the engineassembly 500, a servo and battery package 1404 and steering servos 1202.Of these components, the flanges 302 tilt inwardly the engine assembly500 and the steering servos 1202.

An advantage of the configuration of the chassis 300 is the ability tomount the engine assembly 500 lower with respect to the transmissionassembly 520. Preferably, the transmission assembly 520 is centrallymounted on the central channel area 304, while the engine assembly 500is mounted to the chassis 300 at a lower point on one or more of theflanges 302. The chassis 300 is configured in this manner to preferablyposition the drive shaft 501 of the engine assembly 500 within the rangeof about 3 mm to 13 mm vertically above (of relative to the ground) thelevel of the transmission output shaft 521. The chassis 300 ispreferably press-formed and cut from a sheet of anodized aluminum. Itwill be apparent that the flanges 302 and a central channel area 304 maybe configured in other the variations and configurations to achieve alower center of gravity overall for the vehicle 1400.

In addition to providing a lower center of gravity for the vehicle 1400,the chassis 300 includes forward and rearward extension plates 310, 312positioned at substantially the same vertical level as the centralchannel area 304. The forward and rearward extension plates 310, 312 arepreferably formed integrally with the upper surface of the centralchannel area 304 and support various components of the front suspension,steering and rear suspension assemblies of a vehicle 1400 at a highervertical level than if those assemblies were secured to the flanges 302.Thus, the chassis 300 maintains desirable ground clearance beneath thesuspension and drive assemblies, while providing a relatively low centerof gravity.

In steering systems, for optimum performance, it is important tomaintain geometric parameters within certain desired ranges. Some ofthese well-known parameters are toe-in, camber, caster and roll center.Toe-in is the angle that the wheels make with respect to a line throughthe centerline the vehicle, when viewed from above.

Camber is the inclination of the wheel, from vertical, as viewed fromthe front of the vehicle. It is usually designed to vary with wheeltravel in order to help keep the tire squarely on the ground. Asdescribed elsewhere in this application, camber is adjustable on thevehicle.

Caster is defined as the inclination, from vertical, of the wheel'ssteering axis as viewed from the side of the vehicle. That is, generallyspeaking, caster is a tilt of the steering axis toward the front or backof the vehicle. Basically viewing from the side of the vehicle, draw aline through the upper and lower ball joint of the axle carrier. Theangle off of vertical is the caster. The caster angle is adjusted bymoving the mounting point of the upper arm (effectively the upper balljoint) generally fore and aft with the spacers on the hinge pin of theupper arm. Adjusting caster changes the steering characteristics of thevehicle.

Roll center is adjusted by moving the inner mounting point of the upperarm up and down. This changes the front view Instant Center (IC) of thesuspension. The IC partially defines the roll center.

“Bump steer” can be defined as undesirable steering (toeing in or toeingout) of the wheel/tire during travel (vertical) of the suspensions,assuming that the steering wheel or actuation mechanism is being heldfixed. Bump steer occurs because the toe change is caused by geometricdifferences in the motion arc of the steering control link (toe controllink) and the suspension arms during bump travel of the suspension.Basically, if the vehicle is going straight and then hits a bump with awheel, the raising of the wheel due to the bump changes the toe, causingthe vehicle to tend to veer off without any movement of the steeringwheel/steering actuator. Bump steer tends to be more sensitive to casterand roll center changes than other parameters.

Bump steer is usually impossible to eliminate due to packaging anddesign limitations. Generally, a compromise setting is made to optimallyminimize at the standard suspensions settings. However, having a way toadjust bump steer is desirable due to the range of caster and rollcenter adjustments available in the suspension.

It is known to attempt to minimize bump steer by varying the verticalposition of the mounting points (front view) of the steering controllink on the axle carrier 140′ of the front wheels. Thus, minimizing bumpsteer while adjusting caster and roll center is difficult andcomplicated, requiring extensive trial and error on the part of theoperator. For example, once an adjustment to caster and/or roll centeris made, bump steer is reintroduced by the new settings unless there isa provision for “tuning” it back out.

An embodiment of the present invention incorporates an adjustmentfeature that allows the bump steer to be optimized (minimized) for asubstantially complete set of possible combinations of suspensionsettings; i.e., from 5 degrees to 15 degrees of caster, in 2.5 degreeincrements and for either an “upper” or “lower” roll center position.Referring to FIGS. 53, 54A-E and 55, this is accomplished by providingthe attachment pin of the axle carrier 140′, to which the pivot link 154at the end of the control link is attached, with clearance forpermitting movement of the pivot link 154 up and down on the attachmentpin 1390. Ring-shaped spacers A, B or C, taken from a predetermined setof spacers having predetermined thickness are disposed on the pin 1390above and/or below the pivot link 154 to take up the clearance andposition the pivot link 154 at the optimum position on the pin. Thepredetermined thicknesses for the spacers A, B and C are predeterminedfor each combination of caster and roll center adjustments by geometriccalculations and spacers having the appropriate thicknesses are in akit, along with a table indicating which spacers to use and where toposition them on the pin.

Referring to FIGS. 53, 54A-E and 55, and initially to FIG. 53 thereof, aperspective view of the suspension assembly 1380 for the left frontwheel is depicted. Suspension assembly 1380 includes upper and lowersuspension arms 1382 and 1384, to which is attached an axle carrier140′. Axle carrier 140′ has an arm 1386 having generally vertical pin1390 thereon. Control link 110, which extends from a driven steering arm1242 (not shown) includes a pivot link 154 pivotably attached to pin1390.

FIGS. 54A-E show detailed views of the axle carrier 140′, pin 1390 andpivot link 154 with various predetermined combinations of ring-shapedspacers A-B positioned on the pin, above and/or below the pivot link154. It should be noted that, to replace the spacers, pin 1390 is firstremoved, the spacers and pivot link 154 (or 154″″) placed onto it, andthen the pin is replaced.

In FIG. 53A, a thick spacer of thickness A is disposed above pivot link154 and a thin spacer of thickness B is disposed below the pivot link154. As shown in FIG. 55, this combination is used where there is a 5degree caster and the roll center setting is at the “lower” setting.This combination is also used where there is a 7.5 degree caster and theroll center setting is at the “lower” setting.

In FIG. 54B, a thick spacer of thickness A is disposed above pivot link154 and a thin spacer of thickness B is also disposed above the pivotlink 154. As shown in FIG. 55, this combination is used where there is a5 degree caster and the roll center setting is at the “upper” setting.

In FIG. 54C, a thick spacer of thickness A is disposed below pivot link154 and a thin spacer of thickness B is also disposed below the pivotlink 154. As shown in FIG. 55, this combination is used where there is a10 degree caster and the roll center setting is at the “lower” setting.This combination is also used where there is a 12.5 degree caster andthe roll center setting is at the “upper” setting.

In FIG. 54D, a thick spacer of thickness A is disposed below pivot link154 and a thin spacer of thickness B is disposed above the pivot link154. As shown in FIG. 55, this combination is used where there is a 10degree caster and the roll center setting is at the “lower” setting.This combination is also used where there is a 12.5 degree caster andthe roll center setting is at the “upper” setting.

In FIG. 54E, a “standard” configuration can be employed, where astandard hollow ball pivot link 154″″ is used that has approximatelyequal length collars 155 and 157 at its upper and lower sides that formpart of the pivot link 154″″. Alternatively, spacers can be used thathave the same, medium thickness “C,” thus, positioning the pivot link atthe approximate midpoint of pin 1390. Such a medium positioning islisted in the table of FIG. 55 as “tall center hollow ball.” Thiscentered combination is used where there is a 7.5 degree caster and theroll center setting is at the “lower” setting. This combination is alsoused where there is a 10 degree caster and the roll center setting is atthe “upper” setting.

Of course, because the caster angles and roll center settings will varyby vehicle geometry, weight and other parameters, the above casterangles and roll center settings are only examples for a particularvehicle of a particular geometry, weight and other parameters. Ofcourse, finer increments (such as 1 degree increments for caster andmore increments for the roll center setting) could be employed,resulting in more spacer thicknesses and combinations thereof.

FIGS. 56, 57A through D and 58A through D, illustrate one configurationof a front suspension assembly 1500 secured to a front bulkhead assembly1502 of the vehicle 1400. The suspension assembly 1500 comprises upperand lower suspension arms 1504 and 1506 pivotally mounted to thebulkhead assembly 1502. A rocker arm 1508 is pivotally mounted to a postor boss 1510 extending at an angle into the bulkhead assembly 1502,inboard and above the point of connection of the upper suspension arm1504 to the bulkhead assembly 1502. The rocker arm 1508 is pivotallycoupled to a push rod 1512 and a damper assembly 1514. The outboard endof the push rod 1512 is pivotally secured to the outboard end of thelower suspension arm 1506, urging the suspension arm 1506 outwardly anddownwardly. Upward movement of the suspension arm 1506 displaces thepush rod 1512 inwardly toward the rocker arm 1508, which in turn pivotsto compress the damper 1514 against a pivot pin 1516. Downward movementof the suspension arm 1506 displaces the push rod 1512 outwardly, whichin turn pivots the rocker arm 1508 to release the damper 1514. Therocker arm 1508 is generally triangular in shape. The portion of therocker arm 1508 pivotally connected to the push rod 1512 is referred toas the input arm. A portion of the rocker arm 1508 pivotally connectedto the damper assembly 1514 is referred to as the output arm.

The damper 1514 is generally aligned with the longitudinal axis of thevehicle 1400 and a substantially horizontal position, with a slightupward inclination from the point of connection to the bulkhead assembly1502 toward the point of pivotal connection to the rocker arm 1508. Thesubstantially horizontal position of the damper 1514, mounted adjacentthe points of connection of the suspension arms 1504, 1506 to thebulkhead assembly 1502, reduces vertical space requirements and protectsthe damper 1514 from damage.

The rocker arm 1508 pivots about an axis substantially perpendicular tothe axis of the push rod 1512 at some point during operation of thesuspension assembly 1500. The rocker arm 1508 pivotal axis is orientedto translate movement of the damper assembly 1514 into substantialalignment with the push rod 1512 as the rocker arm 1508 pivots. The pushrod 1512 is mounted to the rocker arm 1508 for pivotal movement alongvertical and horizontal axes relative to the rocker arm 1508. As thesuspension assembly 1500 moves, the push rod 1512 pivots upwardly anddownwardly relative to its point of connection to the rocker arm 1508,following vertical movement of the outboard end of the suspension arm1506.

Referring now to FIGS. 57A through D, the suspension assembly 1500 isshown in the full bump position, with the suspension arms 1504, 1506displaced to their uppermost extent. This position corresponds with thevehicle 1400 reaching a lowermost position relative to an underlyingsurface. In this position, the push rod 1512 rotates the rocker arm 1508toward a damper 1514, substantially fully compressing the damper 1514.

Referring now to FIGS. 58A through D, the suspension assembly 1500 andis shown in the full droop position, with the suspension arms 1504, 1506extended to their lowermost extent. This position corresponds with thevehicle 1400 reaching its highest position relative to an underlyingsurface. In this position, the damper 1514 rotates the rocker arm 1508to fully extend the push rod 1512.

A position intermediate to the full bump and full droop positions is theride height position. In the ride height position, the suspensionassembly 1500 reaches an equilibrium position in which the force exertedby the push rod 1512 counteracts the vehicle weight placed on thesuspension arms 1504, 1506. In general, relative proportions of totaltravel distance of the outboard ends of the suspension arms 1504, 1506at the axle carrier 140′ (i) from ride height to full bump and (ii) fromthe ride height to full droop is referred to as the up/down traveldistribution. The travel distribution of the suspension assembly 1500 isapproximately two-thirds to one third. A ride height of the vehicle 1400can be adjusted by changing the point of connection of the outboard endof the push rod 1512 to the outboard and of the suspension arm 1506.This is accomplished by movement of the push rod 1512 outboard endbetween a number of positioning apertures 1518 to which the push rod issecured by a pin 1520.

The suspension assembly configuration of FIGS. 56 through 64 providesnumerous advantages. Amongst many advantages too numerous to list, butthat will nevertheless be apparent to those skilled in the art, theconfiguration of the suspension assembly 1500 is capable of providingrelatively large motion ratios (MR), a relatively large range of travelbetween full bump and full droop positions, enhanced progressiveness ofthe suspension, as well as the ability to relatively accurately adjustthe suspension progressiveness over the range of movement. The motionratio (MR) is generally described as the ratio of vertical displacementof the wheel to displacement of a corresponding suspension springmember. Depending on the suspension design, motion ratios often varyover the range of suspension travel. Accordingly, it is often useful todefine the motion ratio at various points in the suspension travel. Themotion ratio at a particular point in the travel range is referred to asthe instantaneous motion ratio. A progressive suspension is generallyone in which the suspension spring force at the wheel increasesnon-linearly as the suspension spring member is displaced by verticalwheel travel. Progressiveness can be defined as a change in motion ratio(MR) of the suspension over some range of travel.

Furthermore, a variety of performance characteristics can beindependently adjusted in the assembly 1500, without substantiallyaffecting other performance characteristics. For example, the rideheight of the assembly 1500 can be adjusted without significantlyaffecting the travel distribution or the wheel rate. This is becauseadjustment of the ride height has a relatively insignificant effect on amotion ratio of the suspension assembly 1500.

For example, progression of the suspension assembly 1500 is primarilyaffected by the angle between the input and output arms of the rockerarm 1508, along with the starting angle between the damper 1514 and theoutput arm, as shown by angle A in FIG. 64. The progression rate can berelatively easily varied accurately by substitution of rocker armshaving appropriate dimensions.

As described in pages 42 through 43 of the REVO Owners Manual, appendedhereto as Appendix A and incorporated herein by reference for allpurposes, and on pages 42-43 thereof, the progression rate (orprogressiveness) of the suspension determines the extent to which thespring force produced at the wheel by one or more suspension springmembers being displaced will vary with suspension travel, or verticaltravel of the wheel. A suspension configuration functions progressivelywhen the spring force at the wheel (or suspension force) increases withmovement toward the full bump position, at a progressively increasing,non-linear rate. The non-linearly increasing suspension force of aprogressively functioning suspension can be achieved using one or moreassociated suspension spring members that become progressively stiffer(i.e., the spring rate increases, as does the perceived stiffness of thespring member) with displacement. By comparison, a suspensionconfiguration functions linearly or at constant-rate when the springforce at the wheel (or suspension force) increases with movement towardthe full bump position, at a substantially steadily increasing, linearrate. This linearly increasing suspension force can be achieved usingone or more associated suspension spring members that do not becomesubstantially stiffer with displacement and an associated suspensionassembly linkage that substantially does not function progressively.

It will be apparent to those skilled in the art, that a suspension canbe configured to function progressively through one or more segments ofwheel travel or throughout the entire range of wheel travel. Moreover,the degree of progressiveness can be varied as desired with wheeltravel. The configuration of the suspension and/or variation in thestiffness of the one or more associated spring members can be employedto produce the degree of progressiveness associated with suspensionwheel travel desired.

FIGS. 62A and B and 63A and B illustrate, respectively, rear suspensionassembly and front suspension assembly rocker arms. Variation of thedimensions A, B, C, D and E, as well as the lengths of associatedpushrods will vary the progressiveness of the suspension assemblies.Dimensions associated with a variety of progressiveness and suspensiontravel are listed in Table 1. The dimension values listed in Table 1,except for dimension C (in degrees), can be for millimeters in anembodiment, or for centimeters in another embodiment, or for other unitsof measure in yet other embodiments, depending upon the desired scale orsize of the vehicle. Further, the values presented illustrate therelative proportions of the various components of correspondingembodiments; however, it will be apparent to those skilled in the artthat other dimension values can be substituted, if desired and that thesuspension disclosed is not limited to the dimension values provided.

FIGS. 59 through 61 identify dimensions of the left front and rearsuspension assemblies having motion ratios of approximately 4.5 to 1 andhigh-performance progressiveness curves. The numerical values of thedimensions identified in FIGS. 59 through 61 are shown in Tables 2through 5 below. The dimensions listed in Tables 2 through 5 can be formillimeters in an embodiment, or for centimeters in another embodiment,or for other units of measure in yet other embodiments, depending uponthe desired scale or size of the vehicle. Further, the values presentedillustrate the relative proportions of the various components ofcorresponding embodiments; however, it will be apparent to those skilledin the art that other dimension values can be substituted, if desired,and that the suspension disclosed is not limited to the dimension valuesprovided. Variations of these dimensions will yield various motionratios and progressiveness curves in the suspension assembly 1500. TABLE1 Dimensions of Front and Rear Suspension Assembly Rocker Arms PushrodEnd Rocker Length A B C D E Front Progressive 1 115.55 38.20 20.00 98.008.10 16.20 Progressive 2 120.50 38.40 20.00 88.65 8.10 16.20 Progressive3 125.25 39.45 20.00 80.50 8.10 16.20 Long travel 115.55 40.00 15.2092.50 8.10 16.20 Rear Progressive 1 115.55 30.60 19.00 85.00 3.60 16.70Progressive 2 120.50 30.90 19.00 72.80 3.60 16.70 Progressive 3 125.2532.00 19.00 63.00 3.60 16.70 Long travel 115.55 43.40 19.00 81.00 3.6016.70

Referring now to FIG. 59, values of the dimensions x1-x9 and y1-y8appear in the first part of Tables 2 through 5 below. Table 2 lists thevalues of various dimensions of the suspension utilizing P1 (Progressive1) rocker arms. Table 3 lists the values of various dimensions of thesuspension utilizing P2 (Progressive 2) rocker arms. Table 4 lists thevalues of various dimensions of the suspension utilizing P3 (Progressive3) rocker arms. Table 5 lists the values of various dimensions of thesuspension utilizing LT (Long Travel) rocker arms.

Referring now to FIG. 60, values of dimensions x1-x9 and dimensionsy1-y8 appear in the second part of Tables 2 through 5 below. Table 2lists the values of various dimensions of the suspension utilizing P1(Progressive 1) rocker arms. Table 3 lists the values of variousdimensions of the suspension utilizing P2 (Progressive 2) rocker arms.Table 4 lists the values of various dimensions of the suspensionutilizing P3 (Progressive 3) rocker arms. Table 5 lists the values ofvarious dimensions of the suspension utilizing LT (Long Travel) rockerarms.

Referring now to FIG. 61, values of dimensions x1-x2 and z1-z10 appearin the third part of Tables 2 through 5 below. Table 2 lists the valuesof various dimensions of the suspension utilizing P1 (Progressive 1)rocker arms. Table 3 lists the values of various dimensions of thesuspension utilizing P2 (Progressive 2) rocker arms. Table 4 lists thevalues of dimensions of the suspension utilizing P3 (Progressive 3)rocker arms. Table 5 lists the values of various dimensions of thesuspension utilizing LT (Long Travel) rocker arms. TABLE 2 SuspensionDimensions with P1 Rocker Arms Name Value What Name Value What Frontsuspension, view from front, P1 rocker arms x1 5.5 LCA pivot y1 52.3Lower ball joint/pivot ball x2 12.5 Damper on rocker y2 58.0 Pushrod onLCA x3 26.5 UCA pivot y3 73.0 LCA pivot x4 29.5 Rocker pivot y4 113.3UCA pivot x5 39.9 Pushrod on rocker y5 127.8 Pushrod on rocker x6 131.8Pushrod on LCA y6 127.0 Rocker pivot x7 154.0 Lower ball joint/ y7 137.3Damper on rocker pivot ball x8 165.5 Center of tire y8 97.3 Upper balljoint contact patch x9 153.3 Upper ball joint Rear suspension, view fromrear, P1 rocker arms x1 5.5 LCA pivot y1 52.0 Lower ball joint/pivotball x2 11.8 Damper on rocker y2 50.8 Pushrod on LCA x3 27.1 UCA pivoty3 73.1 LCA pivot x4 30.5 Rocker pivot y4 106.8 UCA pivot x5 33.9Pushrod on rocker y5 118.1 Pushrod on rocker x6 127.8 Pushrod on LCA y6123.5 Rocker pivot x7 155.3 Lower ball joint/ y7 122.8 Damper on rockerpivot ball x8 166.2 Center of tire y8 97.7 Upper ball joint contactpatch x9 154.5 Upper ball joint Top view, P1 rocker arms x1 16.5 FrontDamper z1 90.0 Front Damper Mount Mount x2 11.8 Rear Damper Mount z223.2 Pushrod on Front Rocker z3 16.4 Front Pushrod on LCA z4 11.9 FrontDamper on rocker z5 13.6 Front Rocker pivot z6 88.5 Rear Damper Mount z716.2 Pushrod on Rear Rocker z8 14.7 Rear Pushrod on LCA z9 14.2 RearRocker pivot z10 8.6 Rear Damper on rockerLCA Lower control armUCAUpper control arm

TABLE 3 Suspension Dimensions with P2 Rocker Arms Name Value What NameValue What Front suspension, view from front, P2 rocker arms x1 5.5 LCApivot y1 52.3 Lower ball joint/pivot ball x2 12.6 Damper on rocker y258.0 Pushrod on LCA x3 26.5 UCA pivot y3 73.0 LCA pivot x4 29.5 Rockerpivot y4 113.3 UCA pivot x5 35.7 Pushrod on rocker y5 130.4 Pushrod onrocker x6 131.8 Pushrod on LCA y6 127.0 Rocker pivot x7 154.0 Lower balljoint/ y7 137.3 Damper on rocker pivot ball x8 165.5 Center of tire y897.3 Upper ball joint contact patch x9 153.3 Upper ball joint Rearsuspension, view from rear, P2 rocker arms x1 5.5 LCA pivot y1 52.0Lower ball joint/pivot ball x2 12.8 Damper on rocker y2 50.8 Pushrod onLCA x3 27.1 UCA pivot y3 73.1 LCA pivot x4 30.5 Rocker pivot y4 106.8UCA pivot x5 29.7 Pushrod on rocker y5 120.7 Pushrod on rocker x6 127.8Pushrod on LCA y6 123.5 Rocker pivot x7 155.3 Lower ball joint/ y7 129.1Damper on rocker pivot ball x8 166.2 Center of tire y8 97.7 Upper balljoint contact patch x9 154.5 Upper ball joint Top view, P2 rocker armsx1 16.5 Front Damper z1 90.0 Front Damper Mount Mount x2 11.8 RearDamper Mount z2 24.1 Pushrod on Front Rocker z3 16.4 Front Pushrod onLCA z4 10.9 Front Damper on rocker z5 11.3 Front Rocker pivot z6 88.5Rear Damper Mount z7 17.0 Pushrod on Rear Rocker z8 14.7 Rear Pushrod onLCA z9 14.2 Rear Rocker pivot z10 7.7 Rear Damper on rockerLCA Lower control armUCA Upper control arm

TABLE 4 Suspension Dimensions with P3 Rocker Arms Name Value What NameValue What Front suspension, view from front, P3 rocker arms x1 5.5 LCApivot y1 52.3 Lower ball joint/pivot ball x2 12.7 Damper on rocker y258.0 Pushrod on LCA x3 26.5 UCA pivot y3 73.0 LCA pivot x4 29.5 Rockerpivot y4 113.3 UCA pivot x5 31.8 Pushrod on rocker y5 133.0 Pushrod onrocker x6 131.8 Pushrod on LCA y6 127.0 Rocker pivot x7 154.0 Lower balljoint/ y7 137.4 Damper on rocker pivot ball x8 165.5 Center of tire y897.3 Upper ball joint contact patch x9 153.3 Upper ball joint Rearsuspension, view from rear, P3 rocker arms x1 5.5 LCA pivot y1 52.0Lower ball joint/pivot ball x2 12.9 Damper on rocker y2 50.8 Pushrod onLCA x3 27.1 UCA pivot y3 73.1 LCA pivot x4 30.5 Rocker pivot y4 106.8UCA pivot x5 25.7 Pushrod on rocker y5 123.3 Pushrod on rocker x6 127.8Pushrod on LCA y6 123.5 Rocker pivot x7 155.3 Lower ball joint/ y7 129.0Damper on rocker pivot ball x8 166.2 Center of tire y8 97.7 Upper balljoint contact patch x9 154.5 Upper ball joint Top view, P3 rocker armsx1 16.5 Front Damper z1 90.0 Front Damper Mount Mount x2 11.8 RearDamper Mount z2 25.3 Pushrod on Front Rocker z3 16.4 Front Pushrod onLCA z4 10.9 Front Damper on rocker z5 13.6 Front Rocker pivot z6 88.5Rear Damper Mount z7 17.9 Pushrod on Rear Rocker z8 14.7 Rear Pushrod onLCA z9 14.2 Rear Rocker pivot z10 7.3 Rear Damper on rockerLCA Lower control armUCA Upper control arm

TABLE 5 Suspension Dimensions with LT Rocker Arms Name Value What NameValue What Front suspension, view from front, LT rocker arms x1 5.5 LCApivot y1 52.3 Lower ball joint/pivot ball x2 16.8 Damper on rocker y258.0 Pushrod on LCA x3 26.5 UCA pivot y3 73.0 LCA pivot x4 29.5 Rockerpivot y4 113.3 UCA pivot x5 40.2 Pushrod on rocker y5 128.0 Pushrod onrocker x6 131.8 Pushrod on LCA y6 127.0 Rocker pivot x7 154.0 Lower balljoint/ y7 134.9 Damper on rocker pivot ball x8 165.5 Center of tire y897.3 Upper ball joint contact patch x9 153.3 Upper ball joint Rearsuspension, view from rear, LT rocker arms x1 5.5 LCA pivot y1 52.0Lower ball joint/pivot ball x2 12.7 Damper on rocker y2 50.8 Pushrod onLCA x3 27.1 UCA pivot y3 73.1 LCA pivot x4 30.5 Rocker pivot y4 106.8UCA pivot x5 35.2 Pushrod on rocker y5 118.4 Pushrod on rocker x6 127.8Pushrod on LCA y6 123.5 Rocker pivot x7 155.3 Lower ball joint/ y7 129.1Damper on rocker pivot ball x8 166.2 Center of tire y8 97.7 Upper balljoint contact patch x9 154.5 Upper ball joint Top view, LT rocker armsx1 16.5 Front Damper z1 90.0 Front Damper Mount Mount x2 11.8 RearDamper Mount z2 25.0 Pushrod on Front Rocker z3 16.4 Front Pushrod onLCA z4 10.9 Front Damper on rocker z5 11.0 Front Rocker pivot z6 88.5Rear Damper Mount z7 29.0 Pushrod on Rear Rocker z8 14.7 Rear Pushrod onLCA z9 14.2 Rear Rocker pivot z10 8.0 Rear Damper on rockerLCA Lower control armUCA Upper control arm

Progressiveness can be defined as the change in motion ratio of thesuspension over some range of travel, as described in Appendix C, “RevoSuspension Claims.” Two or more different ranges of travel can beconsidered. Moreover, at each point along any range of travel there isan instantaneous motion ratio (MR). Over a first range of travel, fromfully extended (full droop) to fully compressed (full bump), the changein motion ratio is ΔMR1. Over a second range of travel, from ride heightto fully compressed (full bump), the change in motion ratio is ΔMR2.Additionally, there is an average motion ratio (MR_(ave)), which is theratio of the full range of wheel travel to the full range of damper(including one or more spring members) travel. The average motion ratio(MR_(ave)) is the ratio of vertical displacement of the wheel over itsfull range of travel to displacement of one or more correspondingsuspension spring members (or associated damper) over its entire rangeof travel. It will be apparent to those skilled in the art that ameasure of progressiveness can then be defined as a ratio ofΔMRn/M_(Rave), or the ratio of one change in motion ratio over aparticular range of travel (ΔMRn) to the average motion ratio over anentire range of travel (MR_(ave)), where “n” signifies a particularrange of motion. For example, if ΔMR2 has a value of 0.49 and MR_(ave)has a value of 4.5:1, then the measure of progressivenessΔMR2=0.49/4.5=11%.

Having thus described the present invention by reference to certain ofits preferred embodiments, it is noted that the embodiments disclosedare illustrative rather than limiting in nature and that a wide range ofvariations, modifications, changes, and substitutions are contemplatedin the foregoing disclosure and, in some instances, some features of thepresent invention may be employed without a corresponding use of theother features. Many such variations and modifications may be consideredobvious and desirable by those skilled in the art based upon a review ofthe foregoing description of preferred embodiments. Accordingly, it isappropriate that the appended claims be construed broadly and in amanner consistent with the scope of the invention.

1. An engine mount for a model vehicle engine, comprising: a hingeconfigured to be secured to a vehicle chassis; a vehicle engine having acrank shaft for driving a vehicle, the engine being secured to the hingefor pivotal movement toward and away from a shaft coupled directly orindirectly to a wheel for driving a vehicle.
 2. The engine mount ofclaim 1, further comprising: a first drive gear coupled to the enginefor transmitting drive force to a vehicle transmission; a firsttransmission gear coupled to a vehicle transmission for receiving driveforce from the engine; and wherein the hinge is at least configured topivot the drive gear toward and away from the transmission gear.
 3. Theengine mount of claim 1, wherein the hinge is configured to secure theengine against movement with respect to the chassis other than pivotalmovement about the axis of the hinge.
 4. The engine mount of claim 2,wherein the hinge is configured to secure the engine against movementwith respect to the chassis other than pivotal movement about the axisof the hinge.
 5. The engine mount of claim 2, further comprising: asecond drive gear having a pitch diameter differing from the first drivegear; and wherein the hinge is at least configured to allow the engineposition to be adjusted to substitute the second drive gear for thefirst drive gear and couple the second drive gear to the firsttransmission gear.
 6. The engine mount of claim 2, further comprising: asecond transmission gear having a pitch diameter differing from thefirst transmission gear; and wherein the hinge is at least configured toallow the engine position to be adjusted to substitute the secondtransmission gear for the first transmission gear and couple the secondtransmission gear to the first drive gear.
 7. The engine mount of claim1, wherein the hinge comprises at least one pivot arm that is at leastconfigured to provide greater movement of at least a portion of theengine as the hinge is pivoted.
 8. An engine mount for a model vehicleengine, comprising: a vehicle chassis, wherein the vehicle chassiscomprises a hinged support that is at least configured to support atleast a portion of the engine; and a means for pivoting the hingedsupport with respect to remaining portions of the vehicle chassis. 9.The engine mount of claim 8, wherein the vehicle chassis furthercomprises two support members, wherein the hinged support is coupledbetween the two support members.
 10. The engine mount of claim 9,wherein the hinged support comprises at least one pivot arm that is atleast configured to provide greater movement of at least a portion ofthe engine as the hinged support is pivoted.
 11. The engine mount ofclaim 8, further comprising: a first drive gear coupled to the enginefor transmitting drive force to a vehicle transmission; a firsttransmission gear coupled to a vehicle transmission for receiving driveforce from the engine; and wherein the hinged support is at leastconfigured to pivot the drive gear toward and away from the transmissiongear.
 12. The engine mount of claim 8, wherein the hinged support is atleast configured to secure the engine against movement with respect tothe vehicle chassis other than pivotal movement about the axis of thehinged support.
 13. The engine mount of claim 11, wherein the hingedsupport is at least configured to secure the engine against movementwith respect to the vehicle chassis other than pivotal movement aboutthe axis of the hinged support.
 14. The engine mount of claim 11,further comprising: a second drive gear having a pitch diameterdiffering from the first drive gear; and wherein the hinged support isat least configured to allow the engine position to be adjusted tosubstitute the second drive gear for the first drive gear and couple thesecond drive gear to the first transmission gear.
 15. The engine mountof claim 11, further comprising: a second transmission gear having apitch diameter differing from the first transmission gear; and whereinthe hinged support is at least configured to allow the engine positionto be adjusted to substitute the second transmission gear for the firsttransmission gear and couple the second transmission gear to the firstdrive gear.
 16. An engine mount for supporting a model vehicle engine,comprising: a first support member; a second support member; a hingedsupport member, wherein the hinged support member is coupled between thefirst support member and the second support member; and wherein thehinged support member provides pivotal movement about the axis of thehinged support member.
 17. The engine mount of claim 16, wherein thehinged support member further comprises at least one adjustment slotthat is at least configured to provide a range of pivotal motion of thehinged support member.
 18. The engine mount of claim 17, wherein thehinged support member comprises at least one pivot arm that is at leastconfigured to provide greater movement of at least a portion of theengine as the hinged support member is pivoted.
 19. The engine mount ofclaim 18, wherein the at least one adjustment slot is located at one endof the hinged support member and the at least one pivot arm is locatedat a second end of the hinged support member.
 20. The engine mount ofclaim 19 further comprising: a first fastening means for coupling the atleast one adjustment slot to the first support member and the secondsupport member; and a second fastening means for coupling the at leastone pivot arm to the first support member and the second support member.